1. INTRODUCTION

The Scuderi Split Cycle (SSC) engine concept was first conceived by Carmelo

Scuderi, who filed the first patent for the concept on July 20, 2001 which was granted in

2003. The SSC concept divides the four strokes of a conventional combustion cycle over two

paired cylinders. The first cylinder, referred to as the “compressor” (Figure 1.1), provides

intake and compression strokes. The second cylinder, referred to as the “expander”, provides

power and exhaust strokes. For the SSC prototype engine, the compressor top dead center

(TDC) is phased 20° later than the expander TDC (TDC-e). Throughout this paper, crank

angles are referred to in terms of degrees after TDC of the expander cylinder (ATDC-e). The

SSC architecture can utilize any combination of normally aspirated, turbocharged, spark or

compression ignition combustion strategies.

Figure 1.1. Principal Elements of the Scuderi Split CycleEngine

The two cylinders are connected by a “crossover port”, through which the high

pressure gas is transferred from the compressor cylinder to the expander cylinder between the

compression and power strokes. This crossover (Xovr) event must quickly transfer all the

working fluid in typically 35°crankangle (CA) duration around TDC-e, essentially at

compressor pressure, before combustion can commence. Outwardly opening Xovr valves

(opening away from the pistons) are employed to enable this high mass flow while avoiding

contact between the valves and pistons. The combination of valve opening period and engine

rated speed results in a valve event duration of 1.25 milliseconds (ms). This, together with the

required valve peak lift means very challenging valve acceleration values.

Figure 1.2 compares the Scuderi Xovr valve events to the intake valve lift events of a

standard production gasoline engine and of a Formula 1 engine. The comparison is done by

calculating the pseudo acceleration by dividing the peak lift by the square of the event

duration. To meet these challenging targets required a careful, CAE led design of the

valvetrain and its components, together with experimental validation of the analysis tools and

durability testing of the parts. This paper describes the architecture of the Xovr valvetrain, its

analytical development, and experimental verification.

Figure 1.2. Typical valve operating parameters for various engine types at rated speed

2. EXPERIMENTS

2.1 Definition of Crossover Valvetrain Architecture

The performance modeling of the engine defined the breathing requirements for the

engine, as well as the size of the Xovr valves for both compressor (XovrC valves) and

expander (XovrE valves), their opening timing, opening duration, and peak lift. Table 2.1.1

summarizes the valvetrain functional requirements.

Table 2.1.1 Summary Specification of the Scuderi Xovr Valvetrains

From the start the driving approach to this valvetrain design was to deliver a highly

reliable system which would enable the team to concentrate on combustion development

work on the research engine without concerns for the valvetrain durability.

The need for a high geometrical compression ratio (>30) in both the compressor and

expander cylinders was recognized early on. This, together with the need to time the valve

opening events near TDC, drove the architectural choice to open the valve outwardly, i.e.

away from the piston. This therefore avoids altogether the risk of valve to piston contact as

well as removing the need to machine pockets on the piston top surface, or to recess the valve

seats in the cylinder head. As a result the valve head sealing surface faces inward toward the

piston and the corresponding surface of the valve seat insert faces outward, as shown in

Figure 2.1.1

Figure 2.1.1 Xovr valve actuation elements for initial research engine

The selection of outwardly opening valves led to the adoption of a center pivoted rocker

with a forked end to accommodate the valve stem and engage a valve adaptor from

underneath. The valve adaptor is connected to the valve via standard design collets.

The high valve acceleration rates created by the target lift and periods drove the

requirement for minimum moving masses together with a stiff kinematic chain capable of

transmitting the motion from the cam profile to the valve and to overcome the loads

generated by the high differential pressures on the valve head with the minimum amount of

vibrations. One obvious choice was to use titanium valves. The adoption of a pneumatic

spring in place of a more conventional high strength steel coiled spring achieved both the

objective of very low mass and low vibrations The pneumatic spring piston sits directly on

top of the valve stem, resulting in a direct load path giving the most lightweight solution.

Pneumatic springs also offer the ability to modulate the spring pressure as a function of

engine speed allowing lower pressures to be used at low engine speeds, resulting in reduced

friction.

The requirement for high reliability drove the choice, for the camshaft, rocker, and

valve adaptor, of adopting high strength steel (UTS > 2000 MPa), capable of withstanding

high contact stresses in operation. The rocker and valve adaptor were also coated with a

diamond-like coating (DLC) layer to reduce the dry friction coefficient and improve their

tribological behavior in boundary elasto-hydro-dynamic (EHD) lubrication experienced

during operation.

Another striking feature of this valvetrain in its current design is the large size of the

cam lobe. The base circle radius of the cam lobe is in excess of 40 mm. The combination of

the 2- stroke operation, short valve opening periods, high accelerations and low rocker ratio

(for high stiffness), conspire to the need for the cam in this valvetrain to be larger than

conventional automotive engines. This is mainly because of high contact stresses at the cam

nose due to small radius of curvature, as well as small radii in the concave part of the cam

profile which impose manufacturing challenges and high values of pressure angles. The

“reliability first” approach also led to the adoption of a cam of diameter larger than necessary,

which results in substantial safety margins on lubrication and on the contact stresses the

materials are capable of withstanding. This left open the possibility in the future to design

profiles with more aggressive accelerations without the need for major design changes.

The rocker shaft is mounted eccentrically in its supports to enable lash adjustment. Oil

ports are provided to the rocker shaft to supply pressurized oil to the rocker bore.

2.2 Pressure Induced Loads on the Crossover Valve

Figure 2.2.1 shows that the XovrC valve has to operate with a pressure differential of 16

bar resisting valve opening, and then with 5 bar assisting the valve closing. The pressure

resisting valve opening arises because there is residual pressure in the Xovr port from the last

revolution, and the valve must begin to open before the compressor cylinder reaches Xovr

pressure or it will not be sufficiently open when mass transfer needs to begin. In the case of

the XovrE valve, there is a differential pressure of 27 bar acting resisting the valve opening,

and 9-15 bar (depending on exact valve closing point) resisting the closing. It is the task of

the valve spring, as well as to force the valve to follow the motion imposed by the cam, to

push the valve against its seat and seal it against the excess of compression or combustion

pressure in the cylinders over Xovr port pressures.

The XovrE valve has the shortest opening period, experiences the highest pressure

differential at the beginning of its opening event and needs to be kept shut against the

combustion pressure. The XovrE valvetrain was therefore selected for the experimental

verification.

Figure 2.2.1 Pressure differentials acting on XovrC and XovrE valves 4000RPM full-load

3. ANALYTICAL MODELING

3.1 Valvetrain Dynamic Simulation

To ensure that the performance demands on the Xovr valvetrain would be met, a

lumped mass and stiffness dynamic model of the valvetrain was constructed using

GTVTRAIN software [6]. The model comprised all the moving parts of a single valvetrain as

well as the rocker shaft and camshaft supports. The model was at first populated with

estimated mass and stiffness values and exercised to define the combination which could

achieve the target valve lift and timings whilst maintaining control of the valvetrain

vibrations. This was a very valuable exercise as it drove the decisions on design and

materials, and on the adoption of the pneumatic springs instead of coil springs.

Figure 3.1 shows the simulation results of a version of thevalvetrain using a

conventional progressive coil spring sized to meet the load requirements defined previously

and operating within acceptable static stress range. The natural frequency of such spring was

2000 Hz, very high by normal automotive standards. Despite this the severe accelerations

experienced during the valve opening event induced very high vibration amplitude of the

spring mid coil (also called surge) at speeds greater than 2000 RPM. This in turn resulted in

excessively high fatigue stress ranges in the spring wire, prohibitive for even the best spring

materials and treatments and which would lead to failure. Ultimately the problem with this

type of spring could be summarized in relatively high mass (the mass of the spring

considered was 58 grams) and relatively low internal damping, even when using progressive

coils.

Figure 3.1. Simulation of coil spring dynamics. Top: Valve spring surge amplitude, Bottom:

Maximum valve spring stress amplitude

Pneumatic valve springs on the other hand have intrinsically high progressivity as shown

by the spring frequency chart in Figure 3.2 because the gas pressure in them follows a

substantially adiabatic transformation. This allows tuning the spring progressivity by varying

the spring initial volume. Furthermore the extremely low mass of the DLC coated aluminum

spring piston means that the critical damping of the pneumatic spring is circa 7 times smaller

than the critical damping of the coil spring. The damping in the pneumatic springs is

generated by the friction of the spring piston against the seal. The seal is made of a grade of

filled PTFE, a very low friction material. The sealing action is achieved by allowing the gas

inside the spring to push the seal against the spring piston and this generates enough friction

force to dampen any resonance of the pneumatic spring itself.

Figure 3.2. Pneumatic spring natural frequency. The curve indicates non-linear stiffness due

to the non-linear nature of polytrophic compression

The dynamic model was also used to determine the loads and stresses on components,

help develop the valve lift profile and study the dynamic loads and motions, including the

valve seating behavior. As more detailed information on component mass and stiffness

became available during the design and testing process the dynamic model was updated and

new simulations produced.

The bending stiffness of the rocker was the controlling stiffness of this valvetrain. Its

effect is felt especially during the valve seating portion of the valve opening event.

Simulations were performed to define the value of this stiffness which would be needed to

achieve acceptable valve seating velocities in the whole speed range.

Figure 3.3 shows the effect on the valve seating velocity of a rocker with the design

target stiffness and of another one with 30% lower stiffness. It can be seen that with low

rocker stiffness the valve seating velocity becomes excessive at speeds just above rated (4000

RPM), giving very little margin for error in the overall valvetrain design.

Figure 3.3. Simulation of the effect of rocker stiffness on valve seating velocity

Figure 3.4 shows the relative contribution, for the design which was made and tested,

of the mass and compliance of the individual valvetrain components on the natural frequency

which is in the region of 4000 Hz. The natural frequency is influenced primarily by the

flexibility of the rocker and masses of the rocker and poppet valve. The poppet valve

constitutes the payload, whereas all the other components have the function of controlling the

motion of the valve. The rocker shaft and the camshaft supports are not part of the moving

elements of the valvetrain (although the hole for the rocker shaft does reduce the rocker

bending stiffness) and were therefore sized to add minimum compliance to the overall

valvetrain. Their target stiffness was chosen to be about twice the rocker stiffness. In the

following section the design process for the valve adaptor and for the rocker will be

discussed.

Figure 3.4. Mass and Compliance breakdown contributing to valvetrain natural frequency

3.2 Valve Adaptor Design and Analysis

The design of the valve adaptor focused on two aspects. Then first was to insure that

the closing impacts of the valve head against the valve seat at high engine speeds would not

unseat the collets from the adaptor and the other was the need to make the bottom face of the

adaptor flexible so that it could make flat contact against the forked end of the rocker,

allowing for minor misalignment. The final design is shown in Figure 3.5.

Figure 3.5. Design features in the valve adaptor assembly

The first objective was achieved by using titanium for the collets, a material with a

relatively high friction coefficient, and to use a relatively small 4° cone angle for the collets

themselves. The detailed design of the adaptor assembly tried to favor the capability of the

joint to store strain energy during its assembly by making the titanium collets relatively thick

and by keeping a small, constant wall thickness in the part of the adaptor in contact with the

collects. This was to safeguard against the joint coming loose due to parts relaxing from

general background vibrations or higher than expected impact forces. Furthermore a pressing

operation was included in the assembly procedure where the top of the valve stem was

pressed with a 5000 N load while holding the valve open by the bottom face of the valve

adaptor. This load occurs naturally in operation during the early parts of the valve opening

event at high engine speeds when the rocker imposes high valve accelerations through the

valve adaptor assembly. This in a way makes the design “self-locking”. The pressing

operation however enables presetting of the position of adaptor and collets, which in turns

allows setting the nominal valvetrain lash without having to first run the valvetrain at high

speed.

The second objective was achieved by sizing the part of the adaptor in contact with the

collets, so that its mean diameter was in line with the mean diameter of the bottom surface of

the adaptor touching the forked end of the rocker, and to have a thin connection between the

bottom pad of the adaptor and the part in contact with the collets. This feature allows the

bottom pad to deform in operation and to follow the deformation of the forked rocker

avoiding localized point contact. This final configuration was arrived at with the aid of FEA

simulations modeling the assembly and disassembling loads, as well as simulating the

interaction between the rocker and the adaptor and checking the component for safe cyclic

(fatigue) operation. Figure 3.6 shows how the bottom face of the adaptor follows the rocker

as the operating load is applied.

Figure 3.6. FEA results for valve adaptor assembly and operation: Vertical displacement of

nodes on the bottom face of the adaptor

3.3 Rocker Mass-Stiffness Optimization, Including Simulating the Misalignment As discussed above, the bending stiffness of the rocker under the combined action of

the cam and valve adaptor loads is the controlling stiffness of the valvetrain. In addition, the

rocker inertia is the biggest single contributor to the total equivalent mass at the valve. A

finite element analysis (FEA) investigation was launched to minimize the rocker inertia

without compromising its bending stiffness. The results are presented in Figure 3.7. The

figure shows the target stiffness and inertia as defined by the valvetrain dynamic simulations

as discussed above. It also shows the rocker design at the start of the investigation. At first

material was added onto the rocker in the shape of localized ribs or features in various

positions to identify in which one the effect on increased stiffness was greatest. Once that

answer was found the CAD model of the rocker was modified in a way which further

enhanced their effect and still made the rocker easy to manufacture. After further FEA runs

and modifications the target stiffness and inertia values were achieved. Figure 3.8 shows the

finalized rocker geometry.

Figure 3.7. FEA Analysis: Stiffness-Inertia optimization of rocker

Figure 3.8. Finalized Rocker geometry

Torsional flexibility was designed into the rocker to accommodate a certain degree of

misalignment between the cam lobe and the corresponding pad on the rocker which could

arise in operation due to a combination of manufacturing and assembly tolerances. The

misalignment was simulated in a FEA analysis where the rocker was pressed against a rigid

plane skewed by 0.15° with respect to the rocker axis and compared with the case with no

misalignment. Two levels of forces were used for the analysis: 2500 and 4000 N.

The results are shown in Figure 3.9. The results were assessed comparing the normal

pressure on the rocker surface between the case with and without misalignment. The

increased contact pressure between the two cases did not exceed 10%.

Figure 3.9. Contact analysis showing tolerance of the design to misalignment of rocker shaft and camshaft centerlines. Von Mises stress on surface. Top row: Contact load = 4000 N, Bottom Row: Contact load = 2500 N. Left column: parallel loading, Right column: 0.15degrees misalignment loading

3.4 Cam Profile Design Figure 3.10 shows the two finalized valve lift profiles which were signed off for

testing. The one called “reference” met the lift and duration targets set for the engine XovrE

valve and was adopted as a baseline for the dynamic simulations with early stage components

and later on used for the first round of rig tests. As design progressed and the mass and

stiffness of the components were optimized a new valve lift profile was designed. This new

profile aimed at achieving the highest valve lifts possible within the possibilities of the

valvetrain design while keeping the same duration of the valve opening event. This was made

possible by accurate profiling of the positive acceleration section on the closing side. Both

profiles have constant velocity ramps which are higher than strictly necessary to achieve a

correct theoretical valve seating behavior.

This choice was made so that experimentation on different lash settings could be

carried out during rig testing to determine the most appropriate value of ramp height for the

profiles to be used on the engine. The plot of the valve acceleration shows the resulting radius

of curvature of the cam as procured. These values, albeit small, are feasible with modern day

cam grinding machinery.

Figure 3.10. Details of the two valve lift profiles used on rig testing. Top left: Valve lift,

Bottom left: Valve velocity, Top right: Valve acceleration, Bottom right: Valve jerk.

PERFORMANCE, PREDICTIONS &TESTS

4.1 Test Rig

A test rig was constructed to verify the performance of a single Xovr valvetrain under

the conditions present in the SSC engine including the effect of opening against Xovr port

pressure. The principal motivations for the experimentation were to fully understand the

limitations of the valvetrain prior to performance development of the engine, to verify the

durability of the valvetrain and to verify the analytical valvetrain modeling so that it could be

reliably used to develop cam profiles in future SSC engine performance development work.

A 37 kW Spintron machine was used to provide a suitable testing platform for the valve

dynamics studies. The test “head” was fabricated from aluminum to ensure even temperature

in operation, and has a very stiff design so any structural deformations of the head would be

negligible. The test head houses the valvetrain as designed. The bottom of the valve is visible

from outside the test head, which allows its lift to be measured directly. The test stand has a

three-stage oil pump that was used to supply lubricating oil at pressure to the test rig to

simulate engine conditions. One of these stages is used to scavenge oil from the head, to

better simulate the conditions that will be present in the engine. An oil jet wasinstalled to

lubricate the interface between the rocker, cam and valve adaptor. A 3 kW heater was added

to the oil circuit to enable hot testing. Testing was typically carried out at 90° C. Test head

temperature was monitored by a thermocouple embedded in the structure in close proximity

to the top of the valve guide. A photo of the test rig is shown in Figure 4.1.

Figure 4.1. Xovr valvetrain test rig

The Xovr valve used in the test rig had a threaded section added to its bottom face. This

added negligible weight, and it permitted steel weights to be bolted to the valve as

perturbation masses. These were used to correlate the valvetrain dynamic model. The

locations of the instrumentation in the test head are shown in Figure 4.2. The valve lift was

measured using a Micro Epsilon optoNCDT 1627-4laser. This unit has a 37 kHz bandwidth,

which was marginal but sufficient to capture the valve bounce at high speed. The laser was

mounted on a steel plate cooled by chilled water to isolate it from the temperature of the test

rig. The camshaft angular position was measured using a 720 count optical encoder. A small

volume port was designed into the test head and filled with compressed air at each revolution

of the camshaft while the Xovr valve is shut. This was done to replicate the load the Xovr

valvetrain experiences in the engine due to the differential pressure between the Xovr port

and the expander cylinder. The port configuration was reached by simulating the valve

opening event in the test head with GT-Power. Figure 4.3 shows that the simulated

differential pressure across the Xovr valve for the engine and for the valvetrain rig are

sufficiently close. To minimize the air consumption the Xovr port in the test head was

designed with minimum volume. An adjustable orifice valve was also installed in the supply

line to fine tune the dynamics of the gas flow.

Figure 4.2. Location of instrumentation in the test head of the valvetrain rig

Figure 4.3. Results of simulation of port pressure discharge during valve opening in the valvetrain rig: Differential pressure at XovrE Valve at 4000 RPM

Piezoelectric pressure transducers were used to measure the dynamic pressures in the

pneumatic spring and the Xovr port, while strain gauge type transducers measured the supply

pressures to provide the pegging for the piezoelectric transducers. Bottled compressed air was

used for the pneumatic spring and to pressurize the port to ensure consistent gas properties.

When performing pressurized port testing the valve would open to atmosphere. It was

important for the supplied gas to be dry to eliminate the possibility of icing of the valve seat,

and air was used to avoid filling the test cell with another gas such as nitrogen and creating

the risk of asphyxiation. The static stiffness of the valve and its stem were measured with the

valve closed by adjusting the lash to a set amount, varying the air spring pressure, and

measuring the lash as shown in Figure 4.4.

This was done to compare to calculated values, and any deviation might indicate high

friction in the pneumatic spring or improper seating of the valve. This was repeated at room

temperature, 57° C, and 90° C to ensure that there were no changes in component seating due

to temperature.

Figure 4.4. Measured static valve stiffness

As mentioned above two cam profiles were designed and tested in the rig. The first was

representative of the profile to be used in the preliminary testing of the SSC experimental

engine. The second was intended to drive the valvetrain to its dynamic limits at engine

operating speeds, so that the limitations of the valvetrain dynamics could be explored and the

assumptions made in the analytical modeling verified. The valve lifts measured at low speed

resulting from the two cam profiles with negative lash (i.e. with the valve never touching its

seat) are shown in Figure 4.5.

Figure 4.5. Measured valve lift from two camshaft profiles tested in the rig

The size of the orifice between the pneumatic spring and the supply line has to be

chosen as a compromise between the mean and transient spring dynamics. The plot in Figure

4.6 shows the internal dynamic portion of the pressure in the pneumatic spring over one

revolution at 1500 RPM and 20 bar air supply pressure for various orifice diameters.

Increasing the diameter allows more gas to be forced back into the supply line as the piston

rises and therefore reduced pressure is available to prevent bounce when the valve closes.

After investigating the full speed and pressure range, a 0.2 mm orifice was chosen. The small

orifice diameter keeps the dynamic pressure in the pneumatic spring high, but slows the

response of the filling of the spring from the line. In the case of the SSC, this is tolerable,

because of the relatively short valve event duration compared to a full revolution of the

camshaft.

Figure 4.6. Influence of varying pneumatic spring orifice diameter

The nominal lash for the valve lift profiles described above was 0.10 mm. According to

the dynamic simulations this would leave enough closing ramp to obtain the correct valve

seating behavior. To confirm this, a series of experiments was performed to verify the

dynamic performance of the valvetrain at a variety of lash settings. Figure 4.7 shows the

valve velocity at 0.05, 0.10, and 0.15 mm lash at 4500 RPM and 25 bar pneumatic spring

pressure. This investigation showed that at this high speed, the opening and closing behavior

show further degradation (increased number of bounces and higher seating forces) as the lash

increases above 0.10 mm. If lash falls to less than 0.10 mm, performance will be acceptable

as long as some lash is maintained. Lash less than 0.10 mm may be possible in the event of

wear, manufacturing tolerances, and thermal gradients.

Figure 4.7. Valve velocity varying lash at 4500 RPM at 25 bar pneumatic spring pressure

The valve lift at 0.10 mm lash and 20 bar spring pressure is shown in Figure 4.8 at

1000, 3500, 4000, and 4500 RPM. Also shown are the corresponding dynamic portions of the

pneumatic spring pressure signal measured with the piezoelectric sensor. This signal was

pegged to zero right before the start of the event. These curves illustrate several interesting

aspects of the valve system behavior. The pneumatic spring pressure at low speed shows the

tendency for the pressure to be pushed back through the orifice into the supply line at low

speeds, hence the decrease in the pressure after the lift event. The remainder of the revolution

provides sufficient time for the spring pressure to recover. This pressure reduction is much

smaller as speed increases. The lift curves also illustrate the increase in peak spring force and

the onset of closing bounce as the speed increases.

Figure 4.8. Valve lift measured experimentally at various speeds at 20 bar spring pressure showing onset of closing bounce

In the figure it can be seen that at valve opening the spring pressure starts to rise before

and movement of the valve is detected by the laser and at valve closing spring pressure

continues to drop after the valve has seated. This is because when the valve is seated the

valve spring compresses the valve stem and therefore before the valve can be opened the

compressive deformation on the stem must be relieved. This happens at all speeds; at high

speeds however the inertia of the valve head adds to the phenomenon as can be seen in the

figure by looking at the valve lift traces in the shaft angle interval between 180-190°. At high

speeds, the extra force required to get the valvetrain in motion has created additional elastic

deformation in the valvetrain and as this deformation is released while the valve continues to

open higher peak lifts are recorded, as is evident from the increased pressure and lift at the

higher speeds. This release of elastic deformation occurs as an oscillation, which becomes

evident when the valve velocity and acceleration plots are viewed. As mentioned above, the

valve of the test rig was designed to accept a perturbation mass to provide more correlation

points between the model and the experiments. Figure 4.9 compares the valve lift and

velocity at 3000 RPM, 0.15 mm lash, and 20 bar pneumatic spring pressure without

additional mass, and with 5 g and 10 g additional mass added to the end of the valve. The

additional mass causes the valve to open later for similar reasons to those discussed above for

speed. The change in oscillation due to the additional mass as the elastic energy is released

becomes evident. The increases in mass also increase seating velocity, creating bounce at this

speed where it did not previously exist. In this case, for the valve seating velocity to be

brought down to more acceptable levels the height of the closing ramp would have to be

increased.

Figure 4.9. Valve lift and velocity at 3000 RPM adding perturbation mass

Once the valve dynamics were understood, a portion of the testing was repeated with

the Xovr port pressurized. Xovr port pressure increases the force required to open the valve.

Figure 4.10 shows the results of two tests run at 0.10 mm lash and 20 bar pneumatic spring

pressure at 4500 RPM. One test had the Xovr port pressurized at 20 bar and the other had no

Xovr pressure. At these conditions, the Xovr port pressure delayed the opening of the valve

by ∼1.5° due to the higher force required to lift the valve against the pressure. Once the valve

has been lifted the pressure on both sides of the valve quickly becomes equal and the

remainder of the valve travel is unaffected by the port pressure including the closing bounce

shown. The pneumatic spring traces for the tests with and without port pressure lie on top of

each other, indicating that the pneumatic spring travel was identical. The valve stem therefore

stretched, and the contact pressure between the rocker and the valve adaptor increases with

port pressure as predicted.

Figure 4.10. Effect of 20 bar Xovr port pressure on valve opening at 4500 RPM with 20 bar pneumatic spring pressure

4.2 Comparison of Predicted and Measured Results

For correlation of predicted dynamic and measured results, it was necessary to update

the dynamic model with measured valvetrain stiffness data for the rocker and camshaft

supports, and to include the effect of camshaft bearing clearance on the cam journal

displacement. During this process, it was discovered that some of the deflections in the

system were not fully understood.

Figure 4.11 shows a summary of the total measured and predicted static deflection in

the valvetrain. The set of blue bars on the left of the plot represent the initial attempt at

model correlation. Reasons for the discrepancy were found both in the rig and in the model.

The rocker shaft in the rig was made of hardened steel, and was mounted in bores in the

aluminum test head that were machined to achieve a close fit. As previously described, lash

adjustment is accomplished by placing the rocker bore eccentric with respect to the rocker

shaft centerline, and rotating the rocker shaft. In the test rig, the rocker shaft is clamped at

one end. It was originally assumed that valve actuation forces would push both ends of the

rocker shaft to the bottom of the bores in the test head. The rocker shaft stiffness was

modeled accordingly. In reality it was found that the rocker shaft behaved like a cantilever,

supported only on the clamped side. The rig was therefore modified by splitting the rocker

shaft and allowing it to be supported at both ends, which increased the stiffness and decreased

the measured deflection. This would also better reflect the mounting of the rocker shaft in the

research engine. Measurements on the rig also showed that the initial estimate for the

camshaft support stiffness was 7.5% too high. The valve adaptor assembly stiffness used in

the dynamic model was also slightly too high and was adjusted using the results of the FEA

analysis. After taking these into account, the model correlated with the measured deflection

within 20 μm, which is well within the measurement capability. Understanding the deflection

is important, as the timing of the valve events are so critical in this application, and the static

and dynamic deflections must be known in order to determine the valve lift profile and the

closing ramps needed to achieve the desired valve motion at all speeds.

Figure 4.11. Summary of static stiffness investigation showing total deflection of valve at peak lift as pneumatic spring pressure varied

After these corrections were made to the model, the measured valve lift correlated

well with the simulated dynamic lift (Figure 4.12). The accuracy of the laser measurement

was 60 μm, which is insignificant compared to the experimental scatter which is shown as a

shaded band in the figure. As can be seen the simulated lift profile lies within the

measurement scatter band. The kinematic valve lift profile is also shown for reference.

Figure 4.12. Predicted and measured dynamic valve lifts at 4000 RPM, 20 bar spring fill pressure, 0.15 mm lash Good correlation was achieved between the predicted dynamic and measured valve

velocities (Figure 4.13) and accelerations (Figure 4.14), and it was confirmed that the valve

was seating with the appropriate velocity. What appear to be high frequency vibrations in

these figures is actually noise in the laser signal. As noted above, the total inaccuracy in the

laser measurement was 60 μm, and roughly two-thirds of this was due to a combination of

electrical and atmospheric noise. Changes in air temperature and air currents will affect the

laser signal slightly, and electrical noise in a test cell with 3 kW heaters and 37 kW motors

cannot be entirely avoided without high expense. The 60 μm noise level was deemed

acceptable, and is hardly apparent on the displacement curves. It becomes very obvious

however when the displacement is differentiated to calculate valve velocity and acceleration.

Filtering the signal would have attenuated the valve bounce amplitude, so the signals were

averaged instead. The average of 300 valve lift events are plotted as the mean measurement

lines in the velocity and acceleration traces for comparison with the simulation.

Figure 4.13. Predicted dynamic and measured velocities at 4000 RPM, 20 bar spring fill pressure, 0.15 mm lash

Figure 4.14. Predicted dynamic and accelerations velocities at 4000 RPM, 20 bar spring fill pressure, 0.15 mm lash

4.3 Durability Tests

A 30 hour durability evaluation was run to determine if the Xovr valvetrain was able to

survive initial high-speed engine tests. This test used the more aggressive cam shown in

Figure 4.5 to exert higher lift and seating forces on the valvetrain. The test was run at 4000

RPM to provide reasonably high dynamic forces, so 30 hours of operation equates to 7.2

million cycles. The lash was increased by 0.05 mm from nominal to reduce the height of the

closing ramp therefore increasing the valve

seating velocity to represent worst-case values, the pneumatic spring supply pressure was

held to 20 bar ± 1 bar. Xovr port supply pressure was included and held to 40 bar ± 2 bar, and

the test head temperature was held to 90°C ± 2°C. To obtain the 40 bar pressure difference

across the valve, the cylinder side of the valve was open directly to atmosphere. This

generated considerable noise so a noise abatement cabinet was built around the rig. This also

consumed a lot of air, and sixteen16-packs of bottled compressed air were required to

complete the test. The valve lift, seating velocity, and spring pressure were measured every

fifteen minutes, and the lash and visuals were checked roughly every six hours. Figure 4.15

shows the valve lift and pneumatic spring pressure at 9, 16, 20, 25, and 30 hours into the

durability test. The valve lift did not change measurably during the entire test, although the

spring pressure did increase during the test. When this trend was noticed, the test was paused

and the pneumatic spring was disassembled.

Figure 4.15. Valve lift and spring pressure during durability test

The first pressure increase was noted at 15 hours, and the assembly was disassembled

at 22 hours. The cause for the increased pressure was that oil had leaked past the piston seal,

and this oil reduced the spring volume, therefore increasing its compression ratio. The seal

supplier was contacted and options to address this issue include altering the composition of

the seal material, changing the pressure of the preload spring inside the seal, and changing the

grooving geometry on the seal face. Resolution of this issue is considered part of the normal

development process for a pneumatic spring system.

The remainder of the components showed little to no wear after the test. The DLC

coated parts including the rocker, valve adaptor, and pneumatic piston appeared as new. The

cam journals also appeared as new. There was very light polishing of the valve guide, and the

valve stem seals were in good condition. The lobe of the cam showed uniform light scuffing

and the profilometer trace indicated an average of 4 μm wear at the most severe point on the

lobe. The cam lobe after the test is shown in Figure 4.16.

Figure 4.16. Lobe of camshaft after 30 hour durability test

The only parts with measurable wear were the valve and seat, shown in Figure 4.17.

The wear was uniform around the valve and its sealing action was still intact by the end of the

durability test. The seating surfaces were measured with a profilometer to determine the wear

rates. The wear rate of the valve was 5.2 μm/million cycles, and that of the valve seat was

6.4μm/million cycles. The valve seat in the test rig was made of beryllium copper whereas

the valve seat in the engine will be made of sintered hardened steel alloy (Bleistahl Como 4-

06 FS) which will have a lower wear rate.

Figure 4.17. Valve and seat after 30 hour durability test

CONCLUSION

A fast acting valvetrain has been designed to manage the Xovr gas flow in the SSC

engine. The application requires that a high mass flow of air be transferred from the

compressor cylinder to the expander cylinder over circa 35° crank angle duration. This places

considerable demands on the Xovr valves, which must open rapidly and close without

bounce. The required performance was achieved with the adaptation of existing technology

and by utilizing outward opening valves. The work demonstrated the need for substituting the

traditional coil springs with lighter gas springs as used in racing engines. The geometry and

the tribology of the rocker surfaces was optimized for the high loads experienced and so was

the rocker stiffness-inertia tradeoff. A suite of analysis models was developed and calibrated

to aid in the design of the valvetrain. The dynamic model of the valvetrain built using GT-

Suite was especially instrumental in balancing the requirements for low effective mass and

high stiffness of the moving components, as well as in refining the design of the cam lift

profile and its opening and closing ramps. An experimental rig was built to test the valvetrain

under most of the conditions that it will experience when installed in the engine. This testing

verified the dynamic capability of the Xovr valvetrain to achieve large accelerations and was

also used to calibrate the analytical models. The additional loads on the valvetrain due to the

differential pressure at opening of the Xovr valve were replicated when an extended

durability test of the valvetrain was carried out. The test was completed successfully. The set

of verified analytical models now available will be used to tailor the valvetrain design and

performance as various Scuderi engine configurations are developed in the future

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