CA2315595A1 - Improvements in variable valve timing systems - Google Patents

Abstract

A valve lifter or valve stem that provides a fine contact point with a variable profile cam that can enable its movement across the cam to provide a greater range of valve timing.
The invention also includes a camshaft and camshaft journal system which allows for axial movement of a camshaft based on engine rpm and which further allows for advancing or retarding a camshaft system based on engine rpm.

Description

Improvements in Variable Valve Timing Systems Related application This application is related to Canadian Patent application 2,257,437 which is incorporated herein by reference.
Field of the Invention The invention relates to improvements in variable valve timing for an internal combustion engine.
Specifically, the invention provides a valve lifter or valve stem that provides a fine contact point with a variable profile cam that can enable its movement across the cam in order to provide a greater range of valve timing. In addition, the system provides a camshaft and camshaft journal system which allows for axial movement of a cam shaft on the basis of engine rpm and which allows for advancing or retarding a camshaft system on the basis of engine rpm. Improvements in the materials of the bearing, race, and other surfaces provide non-failing systems.
Background of the Invention The design of an internal combustion engine requires numerous trade-offs between conflicting performance parameters. For example, in the design of an engine, a designer may wish to minimize exhaust emissions and provide increased fuel economy without compromising satisfactory engine performance. In the past, the design of an such an engine would be limited by such conflicting pararmeters leading the designer to compromise between the parameters. As such, designers will often focus on a primary performance goal which may be to the detriment of desired engine performance (for example, torque or idle stability). Such compromises are often caused by the lack of the designers ability to incorporate breathability into the engine, that is an optimal intake of fuel and air and exhaust of spent gases after combustion.
The breathability of an engine is primarily determined by the physical structure of the cam shaft, cam lobes, valve lifters (and the associated push-rods, rocker arms, if applicable). In particular, the physical shapes of the cams and their relative orientation with respect to one another determine the timing of the intake and exhaust valve opening, the duration of opening, and the timing of valve closure which along with the orientation of respective intake and exhaust valves about the camshaft determine the power map of the cylinder. As a result of the working environment and the physical complexity of these components, adjustment during operation of the engine is difficult and accordingly, most engines utilize a fixed cam timing system wherein the relative timing between valve opening and closure does not vary with engine speed. Accordingly, fixed cam timing engines require trade-offs between the performance parameters of the engine.
More specifically, the camshaft function is to open and close valves at the proper time, to fill the cylinders before combustion and to empty them after combustion. The cams are mounted on the camshaft and have a profile which determines the timing of valve opening, the duration of opening and the timing of valve closing. The lifters are in intimate contact with the cam surface and ride the cam surface in order to impart opening/closing forces to the valves. The opening and closing of valves is thereby timed to the rotation of the camshaft which in turn is controlled by the crankshaft.
Accordingly, the physical dimensions or shapes of the cams, lifters and the orientation of the cams with respect to one another are parameters which can be varied in order to obtain desired engine performance.
With respect to the physical dimensions or design of a cam, the following terms are generally used to describe the cam. For example, the base circle of the cam defines the period that the valve is closed, the clearance ramp defines the time of transition between closure and measurable valve lifting, the flank or ramp provides the time for and characteristics of valve opening, the nose defines the time of full valve opening and maximum opening displacement and the duration defines the time that the valve is off its seat.
Each of these parameters of a cam cannot be independently controlled and therefore require compromises between what the physical dimensions of a cam will allow in relation to the other parameters. For example, duration is a compromise between opening the valves long enough to fill and evacuate the cylinders to the loss of dynamic compression by opening the valves too long and increasing lift increases power but is limited by lifter diameter.
With respect to the design of lifters (or tappets), the technology of lifters is variable between engines.
Generally, the primary goals of the design of a lifter is to maintain contact between the lifter surface and cam surface while minimizing noise during operation. There are two classes of lifters, solid and hydraulic with each class providing variable contact ends including flat, mushrooms and rollers. The use of hydraulic lifters generally reduces valve lash and noise. A flat tappet-cam normally has a slight taper across its surface whereas the corresponding tappet end surface is normally marginally convex in order to compensate for mis-aligned lifter bores.
Roller lifters allow for highly aggressive ramp profiles and, as a result, require high valve spring tensions to keep the roller in contact with the cam. Roller lifters also reduce frictional losses between the lifter and cam and thereby will increase the overall power or efficiency of the engine.
Mushroom lifters have a bulge at the end and are used to provide more lift per duration.
The relative orientation of the intake and exhaust cams with respect to one another contributes to defining the power map of the engine. Specifically, the lobe separation angle or overlap determines the time during which the intake and exhaust valves are opened simultaneously, wherein a wider lobe separation angle generally improves idle quality, idle vacuum and top-end power whereas a narrower lobe separation angle decreases idle quality but provides better mid-range torque.
Degreeing a cam is also a parameter which can be used to affect engine performance and refers to altering the point where the cam activates the valves in relation to the crankshaft. Specifically, retarding the cam shaft, that is, opening a valve later relative to the crankshaft moves the power up the rpm band and can increase horsepower while decreasing lower end torque. In contrast, advancing the cam shaft (opening the valves earlier) has the opposite effect.
In order to address some of the problems associated with fixed cam timing, variable cam timing systems have been designed. Generally, such systems provide a cam lobe having a three-dimensional surface and a lifter camshaft which is allowed to move axially over the three-dimensional cam surface.
Accordingly, the axial position of the camshaft will determine the specific cam profile which controls valve timing. Variable valve timing thereby permits the alteration of valve timing during the operation of the engine allowing engine performance to be modified to match operating conditions. Variations in a variable cam system can enable any one of independently phasing the intake cams, independently phasing the exhaust cams, phasing the intake and exhaust equally or phasing the exhaust and intake cams independently of one another.
For example, by diluting the in-cylinder mixture by reducing fuel intake characteristics by providing shorter intake times increases fuel economy but decreases the torque response of the engine. In contrast, by enriching the in-cylinder mixture by increasing fuel intake times by providing more lift and duration leads to an increase in horsepower. A variable valve timing system can accommodate such conflicting objectives by providing different timing profiles depending on the rpm of the engine thereby contributing to improving the breathability of the engine and increasing the manifold pressure.
In high performance applications, the current state-of the-art recognizes the single axis roller or wheel based lifter as the optimal performance enhancing device for valve train operation. However, as the desire for higher engine rpm has grown, it has been found that wheel based lifters will fail under the higher tension springs utilized in the higher rpm engines. Typically, failure occurs in two ways; roller bearing failure in the wheel itself and/or the catastrophic failure of the lifter, both a result of wheel "flat spotting" which produce valve lifter and valve train vibration.
Furthermore, existing wheel-based lifter designs do not provide direct delivery of lubrication to the roller bearing but rather occurs in an indirect way which decreases the ability to dissipate heat from the bearing surfaces. Accordingly, bearing life may be reduced as the wheel may be in direct contact with the bearing race with minimal oil film between the two surfaces.
To achieve maximum bearing life in a single axle based system, the designer must balance three parameters given that the wheel diameter is maximized within the confines of the lifter body. These three factors are roller bearing diameter, axle diameter and wheel thickness.
Each of these parameters must be varied to minimize the compressive and contact stresses on the bearing surfaces, minimize the stresses in the axle and minimize the deflection of the axle which directly affects the contact stresses within the roller bearings.
While past variable valve timing systems have been disclosed, for example in US Patent 2969051, German publication DE 197 55 937, Swiss publication DE 304494 and US Patent 2307926, the lifter/cam contacting systems have experienced neither widespread implementation or success. The reason for this lack of success is postulated to be a result of failures experienced in the actual implementation of such systems. That is, within the harsh operating conditions of an internal combustion engine, it is speculated that previous variable valve timing systems experience bearing failure within the bearings/races of these systems.
Summary of the Invention In accordance with one embodiment of the invention, there is provided, a cam contacting device having a ball bearing wherein the coefficient of thermal expansion of the ball bearing is less than 3x10-6/degree Celcius.
In accordance with another embodiment, there is provided, a cam contacting device having a ball bearing and a bearing race and support wherein the coefficient of thermal expansion of the ball bearing is less than the coefficient of thermal expansion of the bearing race and support.
In accordance with yet another embodiment, there is provided, a cam contacting device having a ball bearing wherein the ball bearing is ceramic.
In accordance with still yet another embodiment, there is provided, a cam contacting device having a ball bearing wherein the ball bearing is any one of silicon nitride or silicon carbide.
In accordance with a further embodiment, there is provided, a cam contacting device wherein the cam contacting surface is selected from any one of Ceralloy 147-31E, 147-31N, 147-lE, or 147-1.
In accordance with a still further embodiment, there is provided, a cam contacting device selected from any one of a radiused wheel, a ball bearing or a semi-spherical surface and wherein the cam contacting surface is selected from any one of Ceralloy 147-31E, 147-31N, 147-lE, or 147-1.
Brief Description of the Drawings These and other features of the invention will be more apparent from the following description in which reference is made to the appended drawings wherein:
FIGURE 1 includes three designs of lifters including a radiused wheel lifter, a sphere based lifter and solid semi-spherical lifter;
FIGURE 2 is a schematic diagram of a combined overhead camshaft system with a spherical bearing valve depressor and valve seat fuel injector;
FIGURE 3 is a schematic diagram of two camshaft systems which can be axially displaced;
FIGURE 4 is a photograph showing the wear patterns of a ceramic bearing lifter and steel bearing lifter;
and, FIGURE 5 is a photograph showing the wear patterns of a steel radiused wheel lifter and ceramic bearing lifter.
Detailed Description of the Invention Figure 1 shows threes designs for lifters in accordance with the invention including a radiused wheel lifter, a sphere based lifter and solid semi-spherical lifter. The end view of a variable valve timing camshaft with a fixed centreline is shown as 1 and a variable valve timing camshaft with a variable centreline is shown as 2.
A radiused wheel lifter 3 is shown with side and end views. In this design, a roller wheel is fixed in the end of the lifter with bearings allowing the roller wheel to rotate about a fixed axis in the lifter base.
A sphere based valve lifter 4 is shown in an assembled view and disassembled with a hydraulic damping system 5, bearing retainer 6 and spherical bearing 7. 35 shows a lubrication port and 36 shows the inner receiving surface of the lifter.
Figure 2 shows a schematic diagram of a combined overhead camshaft system with a spherical bearing valve depressor and valve seat fuel injector. The system includes a variable profile cam 15 in an overhead cam layout, an overhead valve depressor 16 with spherical bearing (or a radiused wheel or half sphere as described above), a cylinder head 17, a valve 18 and valve spring 19. The valve seat 20 may include fuel injector nozzles 21 with fuel delivery line 22. The intake port 23 delivers air to the cylinder through valve 18. The valve depressor 16 may include spherical bearing race 24 seated on a bearing housing 25 and valve spring retainer 26.
Figure 3 shows a schematic diagram of a camshaft system which can be axially displaced. The system includes a variable profile camshaft 27 which may include detachable components as well as a distributor drive gear 32. A splined shaft section 28 is shown with the distributor gear and a rear journal detached 29.

Furthermore, a biasing system is shown to include a spring 31 retained within a detachable journal 33 which receives a splined shaft 34.
Specific designs of the lifter including ceramic bearings selected from silicon nitride and silicon carbide were tested. Properties of the Ceradyne ceralloy family of silicon nitrides are in Table X.
An initial evaluation of the ceramic ball bearing lifter design was tested at Isky Cams Inc. in the Los Angeles county area. The lifter (hereinafter lifter 1) used a ceramic bearing obtained from Ceradyne Inc.
was made of ceralloy 147-31N material. Lifter 1 was initially tested on a spintron as this machine provided the best opportunity for lifter evaluation without the danger of engine damage in the event of lifter failure. The first evaluation was conducted with low pressure valve springs; 200 lbs/inch, to initially determine if the higher contact loads of the lifter/cam surface interface would result in camshaft scoring.
The test cycle was completed in 8 hours and no damage to the lifter or camshaft was evident. A second evaluation was then conducted using a NASCAR (North American Stock Car) specification valve spring with a higher pressure of 8001bs/inch. This test was also conducted over 8 hours and no damage was observed in either the lifter or the cam lobe.
A second phase of evaluation was conducted on a Chevrolet V8 engine; ZZ4 P/N
24502609, under actual running conditions. This test used two different lifter designs, the ball bearing and the radiused wheel and also to evaluate a material variation in the ball bearing design. Figures 4 and 5 show the results of these tests and Table 1 summarizes the lifter/test design.
Table 1 Lifter Design Material Supplier #

1 Ball Bearing lifter Ceralloy 147-31N Lifter body - Shaver with Engines lubrication to race silicon nitride Ball Bearing- Ceradyne 2 Ball bearing lifter #52-100 Alloy Lifter body - Shaver with steel Engines lubrication to race Ball Bearing - Timken 3 Radiused lifter #52-100 Alloy Lifer body and wheel steel - Shaver Engines The test was conducted at the Shaver Engine facility in Torrance California.
The three different designs were placed in the test engine using high performance springs ; p/n 10134358 rated at 3561bs/inch. The engine was started and under load the rpm was controlled at 2000rpm. After 2 minutes, a noticeable miss was detected and the engine operation was suspended. The engine was immediately disassembled and the components inspected.
Figure 4 shows lifter 1 (left) and lifter 2 (right). As shown, lifter 1 has no material loss and has not degraded the camshaft race in any way. Lifter 2 has suffered extensive material loss and has further degraded the cam lobe appreciably. Figure 5 shows lifter 3 (left and centre) and lifter 1 (right). In this case, we see no damage to either the lifter or the camshaft lobe from the test.
Subsequently, the engine was reassembled with lifters 1 and 3 and a new camshaft with the same specification was installed and the testing was continued for 6 hours at various rpm's (idle to 6000) and loads. T'he tested engine produced the same horsepower and torque levels specified and no problems in engine operation were detected. Upon completion, the engine was disassembled and lifters and camshaft were measured for wear. There was no appreciable wear on either the lifters or camshaft lobes.
Comparison of Steel and Ceramic Bearings The failure of the steel ball appears to be from galling (i.e. localized welding) of the steel ball to the steel valve body. Once galling started, the ball would intermittently slide and roll both in the pocket and on the camshaft. This galling and sliding action of the ball would account for its uniform wear (.042"). This unintended sliding action of the ball against the camshaft would cause the severe damage (i.e. groove) to the camshaft as was seen.
The success of the ceramic (Silicon Nitride) ball appears to be a result of the lower coefficient of friction and better heat dissipation properties of the ceramic as described in the comments from others below.
Since the ceramic ball did not gall, it would continue to roll in its pocket and rolling contact with the camshaft would be maintained. This would account for the minimum damage/wear seen on the camshaft.
When the ball rolls on the camshaft, it must slide in the pocket of the lifter body. There is consequently some friction, and heat generation inherent in this design. However, with the lower coefficient of friction of the ceramic, the heat generation would be minimal. The oil supplied through the lifter to the sliding surface between the ball and the lifter body would further reduce this friction as well as cool the ball.
Since the ceramic ball is more rigid than the metal ball, it would not deform as much under load.
Consequently, the heat generated internally in the ball would be less in the ceramic ball.
The steel roller assembly has roller elements in it. Consequently, there would be rolling action of the roller against the camshaft. As was the case with the ceramic ball, the wear on the camshaft would be therefore minimized. Work hardening would occur on the camshaft as a result of the contact stress. This is most likely the cause of the narrow band that was seen on the camshaft for both the roller and ceramic ball setup. Since this setup is all rolling action and no sliding action, the friction, and consequently heat generation, would be minimal.
For the silicon nitride ball:
a) The total friction on the ball is less than that of a steel ball.
b) The lower level of friction will generate less heat at the ball contact surfaces than a steel ball.
c)Any heat generated at the ball-cam and the ball-cup (36) interfaces will find its primary dissipation path through the steel interfaces rather than the ball as the thermal conductivity of the ball is significantly lower than that of the steel with the oil supplied to the ball/lifter providing further cooling.
d) The steel contact surfaces will therefore heat up more so than the ball.
e) The steel contact surfaces will therefore expand more than the ball due to the increased heat and the higher coefficient of expansion.
f) Given the different expansion coefficients, the ball will always remain smaller than the surrounding cup. Therefore, the ball should not seize due to heat buildup.
g) The mark on the cam was probably a result of the contact interface heat effectively work hardening the cam.
For the steel ball:
a) The total friction on the ball will be greater than that of silicon nitride ball.
b) The higher level of friction will generate more heat at the ball contact surfaces than the silicon nitride ball.
c) Any heat generated at the ball-cam and the ball-cup (36) interfaces will dissipate through the ball as well as the contact surfaces with the oil supplied to the ball/lifter interface providing further cooling.

d) The steel contact surfaces with therefore heat up as the ball heats up.
e) The steel ball that contacts both the cam and the cup could heat up faster than the cup due to contact with the cam. If the ball heats up faster than the cup, the ball would expand faster than the cup increasing friction and possibly start to seize in the cup.
f) Given the same expansion coefficients, the ball may seize in the cup if the ball heats up faster than the cup.
g) The wear on the cam, ball, and cup was probably a result of the ball starting to seize (gall) in the cup effectively increasing friction and wear on all three surfaces.
For the steel (radiused )wheel:
a) The total friction on the steel wheel will be greater than that of silicon nitride ball.
b) The higher level of friction will generate more heat at the wheel contact surfaces than the silicon nitride ball.
c) Any heat generated at the wheel-cam and the wheel-roller (4) interfaces will dissipate through the wheel rim as well as the contact surfaces.
d) The steel contact surfaces with therefore heat up as the ball heats up.
e) The steel wheel has a smaller contact point than the silicon nitride ball and will generate more heat for a given spring load than the ball.
f) The steel wheel has a large amount of clearance to the supporting lifter surfaces.
g) The wheel, under thermal expansion, would not contact any of the surrounding lifter surfaces therefore the steel wheel should not seize (gall) as the steel ball did.
h) Any heat generated in the wheel rim can only be dissipated through the cam and the roller bearings; there is no forced oil-cooling bath.
i) The steel wheel will get hotter than the silicon nitride ball due to higher levels of friction and no oil both.
j) The mark on the cam was probably due to the high level of heat transferred at the wheel-cam interface. The cam was being effectively work hardening by the heat of friction and contact pressure. The mark was more extensive than the silicon nitride ball due to higher temperature and pressure at the cam contact point.
In conclusion, the success of the Silicon Nitride may be a result of a hardening of the Cam lobe Metal as a result of the heat generated by the contact of Ceramic and Metal as well as the thermal conductivity and thermal expansion coefficient differences.
As follows are some property variations between steel and silicon nitride which likely account for the success of one over another.
Property Alloy Value Thermal Conductivity 147-31N (silicon nitride)20-30 W/mK

1020-1080 45 to 50 W/mK

Coefficient of thermalsilicon nitride 3x10 -6/C
expansion bearing steels ~ 9-2x10-6/C

RCF (rolling contact higher in silicon nitride fatigue) than bearing steels Ceradyne's Ceralloy Family of Silicon Nitrides Silicon Nitride Properties Property 3 E lloy 147- Ceralloy 147-31 N Ceralloy 147-1 E Ceralloy 147-1 Process ~ Sinter Sinter ~ Hot Press Reaction Bonded Density (g/cc) 3.25 3.2t ~ 3.1 2.4 Density (% gg..3 >gg,5 >98.5 75 Theoretical) Flexural Strength 700 ~ 800 700 240 (MPa) cL'0 RT
Weibull Modulus 10-15 15-30 18 10 Elastic Modulus 310 310 310 175 (GPa) Poisson's Ratio 0.27 0.27 0.27 0.22 Hardness HV 1800 1800 1800 800 (0.3) Kg/mm2 Fracture Toughness (MPa 6.0 5.8 5.0 2.5 m1/2) Abrasive Wear Resistance 1130 1110 1120 360 Parameter "' Thermal Expansion Coeff. 3.1 3.1 3.2 3.2 10-6/C; (RT -1000 C) Thermal Conductivity 26 26 42 14 (W/mK) 25 C
Thermal Shock 530 ~ 610 540 330 Parameter (C) Elecrical - _ Resistivity (ohm- 10~14 10~14 10"14 10"14 cm) Dielectric 9 g 9 Constant Alkaline Weight Loss (mgfin2) 2.5 30% NaOH ~
82C - 48hrs Acid Weight Loss (mg/in2) 50°~
H2S04 ~ 82C - 5.0 48hrs

Claims (8)

1. A cam contacting device having improved thermal dissipation properties.

2. A cam contacting device having a ball bearing wherein the coefficient of thermal expansion of the ball bearing is less than 3x10 -6/degree Celcius.

3. A cam contacting device having a ball bearing and a bearing race and support wherein the coefficient of thermal expansion of the ball bearing is less than the coefficient of thermal expansion of the bearing race and support.

4. A cam contacting device having a ball bearing wherein the ball bearing is ceramic.

5. A cam contacting device having a ball bearing wherein the ball bearing is any one of silicon nitride or silicon carbide.

6. A cam contacting device wherein the cam contacting surface is selected from any one of Ceralloy 147-31E, 147-31N, 147-1E, or 147-1.

7. A cam contacting device selected from any one of a radiused wheel, a ball bearing or a semi-spherical surface and wherein the cam contacting surface is selected from any one of Ceralloy 147-31E, 147-31N, 147-1E, or 147-1.

8. A cam contacting device as in any one of claims 1-7 wherein the cam/cam contacting device interface is lubricated from within the cam contacting device.