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Evaporative Emission Control EVAP

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In einem Thread über Chevrolet Caprice kam die Frage nach dem EVAP System, seine Funktion und Diagnose auf. Hier einmal mehr über das System, zuerst der allgemeine Wikipedia Eintrag dazu:

Zitat:

Evaporative emissions are the result of gasoline vapors escaping from the vehicle's fuel system. Since 1971, all U.S. vehicles have had fully sealed fuel systems that do not vent directly to the atmosphere; mandates for systems of this type appeared contemporaneously in other jurisdictions. In a typical system, vapors from the fuel tank and carburetor bowl vent (on carbureted vehicles) are ducted to canisters containing activated carbon. The vapors are adsorbed within the canister, and during certain engine operational modes fresh air is drawn through the canister, pulling the vapor into the engine, where it burns.

Hier zB die System Beschreibung für ein "moderneres" Fahrzeug

Document ID# 304317 - 1999 Chevrolet/Geo Alero Export
Evaporative Emission (EVAP) Control System Operation Description

Purpose

The basic Evaporative Emission (EVAP) Control system used on all vehicles is the charcoal canister storage method. This method transfers fuel vapor from the fuel tank to an activated carbon (charcoal) storage device (canister) to hold the vapors when the vehicle is not operating. When the engine is running, the fuel vapor is purged from the carbon element by intake air flow and consumed in the normal combustion process.

Enhanced EVAP System Operation

FIGURE EVAP Control System Overview(c)
BILD
(1) EVAP Vent Valve/Solenoid
(2) EVAP Vent Valve/Solenoid Ignition Feed Circuit Terminal
(3) EVAP Vent Valve/Solenoid Control Circuit Terminal
(4) EVAP Vent Valve/Solenoid Filter
(5) EVAP Vapor Lines
(6) Fuel Tank Pressure Sensor
(7) Fuel Tank Pressure Sensor Ground Circuit Terminal
(8) Fuel Tank Pressure Sensor Signal Circuit Terminal
(9) Fuel Tank Pressure Sensor Circuit 5 Volt Reference Circuit Terminal
(10) Fuel Filler Pipe
(11) Check Valve (Spitback)
(12) Modular Fuel Sender Assembly
(13) Fuel Limiter Vent Valve (FLVV)
(14) Pressure/Vacuum Relief Valve (Optional)
(15) EVAP Canister
(16) EVAP Canister Purge Valve/Solenoid Ignition Feed Circuit Terminal
(17) EVAP Canister Purge Valve/Solenoid Control Circuit Terminal
(18) Intake Manifold Vacuum Source
(19) EVAP Canister Purge Valve/Solenoid
(20) EVAP Service Port
The EVAP purge solenoid valve allows manifold vacuum to purge the canister. The Powertrain Control Module (PCM) supplies a ground to energize the EVAP purge solenoid valve (purge on). The EVAP purge solenoid control is Pulse Width Modulated (PWM) or turned on and off several times a second. The PCM controlled PWM output is commanded when the appropriate conditions have been met:

• Engine coolant temperature above 25°C (77°F).
• After the engine has been running about 2.5 minutes on a cold start or 30 seconds on a warm start.
• The vehicle is operating in closed loop fuel control.

Canister purge PWM duty cycle varies according to operating conditions determined by mass air flow, fuel trim, and intake air temperature. Canister purge will be disabled if TP angle increases to above 70%. Canister purge will be re-enabled when TP angle decreases below 66%.

The evaporative leak detection diagnostic strategy is based on applying vacuum to the EVAP system and monitoring vacuum decay.

The fuel level sensor input to the PCM is used to determine if the fuel level in the tank is correct to run the EVAP diagnostic tests. To ensure sufficient volume in the tank to begin the various diagnostic tests, the fuel level must be between 15% and 85%.

The PCM monitors system vacuum level via the fuel tank pressure sensor input.

OBD II Evaporative Emission System - Fuel Tank Vacuum Sensing General Description (Enhanced EVAP)
The evaporative system includes the following components:

• Fuel tank
• Evaporative emission canister vent solenoid
• Fuel tank pressure sensor
• Fuel pipes and hoses
• Vapor lines
• Fuel cap
• Evaporative emission canister
• Purge lines
• Purge valve solenoid

Results of Incorrect Operation

• Poor idle, stalling and poor driveability can be caused by:
• Malfunctioning purge solenoid.
• Damaged canister.
• Hoses/lines split, cracked and/or not connected properly.

EVAP Purge Solenoid, EVAP Vent Solenoid and Fuel Tank Pressure Sensor
The evaporative leak detection diagnostic strategy is based on applying vacuum to the EVAP system and monitoring vacuum decay.

Before the EVAP system diagnostic tests are run the following conditions must be present:

• No TP sensor, ODM, IAT sensor, or MAP sensor DTCs set.
• Engine coolant temperature is between 4°C and 30°C (40°F and 86°F).
• Start up engine coolant temperature is not more than 8°C (14°F) greater than start up intake air temperature.
• Intake air temperature is between 4°C and 30°C (40°F and 86°F).
• Start up intake air temperature is not more than 2°C (4°F) greater than start up engine coolant temperature.
• Fuel tank level is between 15% and 85%
• BARO is above 75 kPa.

The EVAP system diagnostic tests will be run following a cold start, as indicated by the ECT and IAT sensors. The fuel level sensor input to the PCM is used to determine if the fuel level in the tank is correct to run the EVAP diagnostic tests. To ensure sufficient volume in the tank to begin the various diagnostic tests, the fuel level must be between 15% and 85%.

The PCM monitors vacuum level via the fuel tank pressure sensor input. At an appropriate time, the EVAP purge solenoid and the EVAP vent solenoid are turned on, allowing engine vacuum to draw a small vacuum on the entire evaporative emission system. After the desired vacuum level has been achieved, the EVAP purge solenoid is turned off, sealing the system. A leak is detected by monitoring for a decrease in vacuum level over a given time period, all other variables remaining constant. A small leak in the system will cause DTC P0442 to be set.

If the desired vacuum level cannot be achieved in the test described above, a large leak or a faulty EVAP purge solenoid is indicated. This can be caused by the following conditions:

• Disconnected or faulty fuel tank pressure sensor
• Missing or faulty fuel cap
• Disconnected, damaged, pinched, or blocked EVAP purge line
• Disconnected or damaged EVAP vent hose
• Disconnected, damaged, pinched, or blocked fuel tank vapor line
• Disconnected or faulty EVAP canister solenoid
• Disconnected or faulty EVAP vent solenoid
• Open ignition feed circuit to the EVAP vent or purge solenoid
• Damaged EVAP canister

Any of the above conditions can set DTC P0440.


A restricted or blocked EVAP canister vent path is detected by drawing vacuum into the EVAP system, turning off the EVAP vent solenoid and the EVAP purge solenoid (EVAP vent solenoid Open, EVAP purge PWM 0%) and monitoring the fuel tank pressure sensor input. With the EVAP vent solenoid open, any vacuum in the system should decrease quickly unless the vent path is blocked. A blockage can be caused by the following conditions:

• Faulty EVAP vent solenoid (stuck closed)
• Plugged kinked or pinched vent hose
• Shorted EVAP vent solenoid driver circuit
• Plugged evaporative canister.

If any of these conditions are present, DTC P0446 will set.

The system checks for conditions that cause the EVAP system to purge continuously by commanding the EVAP vent solenoid on and the EVAP purge solenoid off (EVAP vent solenoid CLOSED, EVAP purge PWM 0%). If fuel tank pressure level increases during the test, a continuous purge flow condition is indicated. This can be caused by the following conditions:

• EVAP purge solenoid leaking
• EVAP purge and engine vacuum source lines switched at the EVAP purge solenoid
• EVAP purge solenoid driver circuit grounded

If any of these conditions are present, DTC P1441 will set.

Refer to the DTC charts for further diagnostic procedures regarding the EVAP system.

Visual Check of Evaporative Emission Canister
Cracked or damaged, replace canister.
Fuel leaking from the canister, replace canister and check lines and line routing.

Hier nur die Beschreibung des älteren Systems, anhand eines 1980 Chevrolet Caprice

Fig. 1: The evaporative emission canister is usually mounted in the side of the engine compartment
fig01

Fig. 2: Evaporative emission control system schematic for early model "open" design canisters
fig02

Fig. 3: Common evaporative emission control schematic for Chevrolet 5.0L engines equipped with thermal vacuum control
fig03

Fig. 4: Common closed canister evaporative control schematic for carbureted vehicles equipped with a vacuum solenoid
fig04

Fig. 5: Evaporative emission control schematic — fuel injected vehicles
fig05

All gasoline vehicles covered by this manual are equipped with an evaporative emission control system which is designed to reduce the amount gasoline vapors which escape into the atmosphere. Float bowl emissions are controlled by internal carburetor modifications and, on later model vehicles, by a vapor line to the canister. Redesigned bowl vents, reduced bowl capacity, heat shields and improved intake manifold-to-carburetor (or throttle body on fuel injected vehicles) insulation reduce vapor loss. The venting of fuel tank vapors into the air has been stopped by means of the carbon canister storage method. This method transfers fuel vapors to an activated carbon storage device which absorbs and stores the vapor that is emitted from the engine's induction system while the engine is not running. When the engine is running, the stored vapor is purged from the carbon storage device by the intake air flow and then consumed in the normal combustion process. The system, in its simplest form, works when manifold vacuum reaches a certain point and opens a purge control valve mounted atop or near the charcoal storage canister. This allows air to be drawn into the canister, thus forcing the existing fuel vapors back into the engine to be burned normally.

The purge function on most earlier model vehicles was controlled by a Thermal Vacuum Switch (TVS) located inline between the canister and the carburetor/intake manifold. The thermal vacuum switch was threaded into a coolant passage such as the thermostat housing and would be activated by engine coolant temperature. As the engine warmed, the switch would open to allow vacuum to the canister or canister control valve.

Later vehicles switched from thermal to electronic control. The purge control on the 231 and on all fuel injected engines is electronically controlled by a normally opened inline purge solenoid which is itself activated by the Electronic Control Module (ECM). On the 231 and most fuel injection systems through 1987, when the system is in the Open Loop mode, the solenoid valve is energized, blocking all vacuum to the purge valve. When the system is in the Closed Loop mode, the solenoid is de-energized, thus allowing existing vacuum to operate the purge valve. This releases the trapped fuel vapor and it is forced into the induction system.

For almost all 1988–89 fuel injected vehicles, a new purge control solenoid was used. This solenoid was a Normally Closed (N/C) component which worked on opposite signals to its predecessor. On these vehicles the ECM would de-energize the solenoid during cold engine operation or idle conditions. When de-energized the solenoid would block all vacuum preventing canister purge. Once the engine was fully warmed and operated above idle, the ECM would energize the solenoid, allowing vacuum to purge the canister of stored fuel vapors.

Many of the carbureted vehicles covered by this book are equipped with a float bowl vent to the canister. On these vehicles a vacuum valve is used to prevent vapor purge from the float bowl when the engine is running. Whenever the engine is off, the valve allows vapors to travel from the float bowl to the canister.

Though a few of the earlier vehicles covered here were equipped with carbon canisters of the "Open" design, meaning that air is drawn in through the bottom (filter) of the canister, most are equipped with a "Closed" design canister which uses a sealed bottom. On later model vehicles equipped with "Closed" design canisters, incoming air which is drawn directly from the air cleaner.

SERVICE

Besides a periodic visual inspection of the system's components, the only periodic service necessary (on early model vehicles so equipped) is canister filter replacement. Later vehicles are equipped with a sealed canister that is not equipped with a replaceable cartridge. On these vehicles, the entire canister assembly must be replaced if any damage occurs or any problems are found with the canister itself.

NOTE: Remember that the fuel tank filler cap is an integral part of the system in that it is designed to seal in fuel vapors. If it is lost or damaged, make sure the replacement is of the correct size and fit so a proper seal can be obtained.

Periodically check for cracks or leaks in the vacuum lines or in the canister itself. The lines and fittings can usually be reached without removing the canister. Cracks or leaks in the system may cause poor idle, stalling, poor driveability, fuel loss or a fuel vapor odor.

Vapor odor and fuel loss may also be caused by; fuel leaking from the lines, tank or injectors, loose, disconnected or kinked lines or an improperly seated air cleaner and gasket.

If the system passes the visual inspection and a problem is still suspected, check the basic operation of the components:

The line from the fuel tank to the canister must be clear and unobstructed. When the engine is OFF, air should pass from the fuel tank towards the canister freely in order to allow vapors to collect in the canister. Make sure the line is free of kinks or obstructions. While the engine is not running, air should NOT be allowed out of the canister.
If equipped with a float bowl vent, the vacuum valve should only allow air to be blown from the carburetor float bowl towards the canister when no vacuum is applied (engine is not running). To test this valve, attempt to blow air through the valve towards the canister with the engine OFF, there should be little or no restriction. Use a hand vacuum pump to apply approximately 15 in. Hg (51 kPa) to the valve, now air should no longer flow towards the canister.
Thermal valves are usually designed to open, allowing vacuum or air pressure towards the canister or control valve only when the engine is warm. Attach a length of hose to the engine side fitting and try blowing towards the canister. Air should be felt at the canister side fitting, only when the engine is at normal operating temperature.
NOTE: When testing valves by blowing air through them, be careful that you are blowing in the proper direction of flow. Many valves are designed to only allow air to flow in one direction and a proper working valve may seem defective if it is tested with air flow only in the wrong direction.

Normally open solenoid valves, which are used on some engines before 1988, should close when energized and open when de-energized. Try blowing air through the valve fittings when the engine is OFF, it should flow with little or no resistance. When the engine is running the solenoid should energize during engine warm-up and de-energize once it has reached normal operating temperature.
Normally closed solenoid valves, used on most 1988–89 engines, should open when energized and close when deenergized. Try blowing air through the valve fittings when the engine is OFF, air should not flow. When the engine is running the solenoid should de-energize during engine warm-up and energize once it has reached normal operating temperature.
Most vacuum and control valves, with the exception of the float bowl valve, are designed to open when vacuum is applied. In either case, all are designed to allow air to pass through only during one condition (vacuum on or off depending on design). To test vacuum valves, try to blow air through the valve with and without vacuum applied. If air can pass through during only 1 of these conditions the valve is likely operating properly. On the other hand if air can always or never flow, the valve is defective. Use a hand vacuum pump to apply approximately 15 in. Hg (51 kPa) to the valve. The valve should open or close (as applicable) and hold the vacuum for at least 20 seconds, or the diaphragm is leaking and the valve must be replaced.

Und nochmal etwas offizieller für einen Document ID# 39915 - 1996 Chevrolet/Geo Caprice
Evaporative Emission (EVAP) Control System

The Evaporative Emission (EVAP) control system used on all vehicles is the charcoal canister storage method. This method transfers fuel vapor from the fuel tank to an activated carbon (charcoal) storage device (canister) to hold the vapors when the vehicle is not operating. When the engine is running, the fuel vapor is purged from the carbon element by intake air flow and consumed in the normal combustion process.

Fuel Vapor Canister
fig06
(1) Tank Tube
(2) Air Tube (Fresh Air Inlet)
(3) Purge Tube
The Evaporative Emission (EVAP) control system uses a 1500 cc charcoal canister to absorb fuel vapors from the gas tank.

When gasoline vapor builds enough to overcome the spring tension of the EVAP pressure control valve, the vapor will flow to the canister where it is absorbed and stored by the charcoal. Under certain operating conditions the PCM will command the purge solenoid valve to open. This allows the vapor to flow into the intake manifold for combustion.

This system has a remote mounted canister purge control solenoid valve. The PCM operates this solenoid valve to control vacuum to the canister. Under cold engine or idle conditions, the solenoid valve is closed, which prevents vacuum from being applied to the canister. The PCM activates (or opens) the solenoid valve and allows purge under the following conditions:

• Engine is warm.
• After the engine has been running a specified period of time.
• Above a specified road speed.
• Above a specified throttle opening.

EVAP Pressure Control Valve
fig07
(1) Control Tube
(2) Tube to Canister
(3) Umbrella Valve
(4) Restriction
(5) Tube to Fuel Tank
(6) Diaphragm
(7) Diaphragm Spring

Evaporative Emission (EVAP) Pressure Control Valve
This system uses an in-line EVAP pressure control valve as a pressure relief valve. When vapor pressure in the tank exceeds approximately 5 kPa (.7 psi) the diaphragm valve opens, allowing vapors to vent to the canister. A 1.14 mm (0.045 inch) orifice in the passage leading to the canister tube causes pressure to drop slowly, preventing the valve from oscillating (buzzing). When the tank pressure drops below 5 kPa (.7 psi), the valve closes causing vapors to be held in the fuel tank.
Results of Incorrect Operation

• Poor idle, stalling and poor driveability can be caused by the following:

1. Inoperative purge solenoid valve.
2. Damaged canister.
3. Hoses split, cracked and, or not connected to the proper tubes.

• Evidence of fuel loss or fuel vapor odor can be caused by:

1. Liquid fuel leaking from fuel lines.
2. Cracked or damaged canister.
3. Inoperative canister control valve.
4. Vacuum hoses that are:

• Disconnected.
• Mis-routed-routed.
• Kinked.
• Deteriorated or damaged vapor hoses.

If the solenoid valve is open, or is not receiving power, the canister can purge to the intake manifold at the incorrect time. This can allow extra fuel during warm-up, which can cause rough or unstable idle.

EVAP Vacuum Switch
fig08
EVAP Purge Vacuum Switch
The EVAP Purge Vacuum Switch is used by the PCM to monitor EVAP canister purge solenoid operation and purge system integrity. The EVAP Purge Vacuum Switch should be closed to ground with no vacuum present (0% EVAP Purge PWM). With EVAP Purge PWM at 25% or greater, the EVAP Purge Vacuum Switch should open.

An incorrect EVAP Purge system flow should set a DTC P0441. A continuous purge condition with no purge commanded by the PCM should set a DTC P1441. Refer to Evaporative Emission (EVAP) Control System for a complete description of the EVAP system.

The Evaporative Emission (EVAP) canister purge solenoid valve and the EVAP Vacuum switch diagnosis is covered in the following DTCs:

• P0441
• P0443
• P1441

EVAP Control System Schematic
fig09
(1) Throttle Body
(2) EVAP Solenoid Valve
(3) EVAP Vacuum Switch
(4) EVAP Canister
(5) EVAP Pressure Control Valve
(6) Floating Roll-Over Valve
(7) Fuel Tank
Visual Check of EVAP Canister
If cracked or damaged, replace EVAP canister.

Evaporative Emission (EVAP) Pressure Control Valve

With a hand vacuum pump, apply approximately 51 kPa (15 in. Hg) to the control vacuum tube. After ten seconds, there should be at least 17 kPa (5 in. Hg) vacuum remaining. Be sure the hand vacuum pump being used does not have an internal leak and the hose connections to control vacuum tube and pump are secure. If after 10 seconds there is less than 17 kPa (5 in. Hg) vacuum, the valve must be replaced.

With 51 kPa (15 in. Hg) vacuum still applied to the control vacuum tube, attach a short piece of hose to the valves tank tube side. Blow into the tube. You should feel the air pass through the valve. If air does not pass through, the valve must be replaced.

Automatikgetriebe Abmessungen

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transmission-dimensions

Abmessungen gebräuchlicher US Automatikgetriebe

Emission testing

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EMISSIONS TESTING
Exhaust all the options

In emissions testing, technology is king. It plays the leading role in ensuring legislation targets are achieved, while enabling engineers to face increasingly difficult challenges.
One of the leading suppliers of systems and equipment facing these challenges is Horiba Automotive Test Systems. Les Hill, leader of the
company’s global product planning group, says, “Nowadays, having a successful test facility is not just a simple question of accurate emissions measurement capability. You also need an accurate and repeatable engine or chassis dynamometer to simulate the loading during operation, the test cell environmental conditions and their control, and the test automation system. Only when you put all these together can you produce data that’s realistic, repeatable and reproducible.”
Also vital for precise emissions measurement is the means by which gas from the exhaust is extracted, conditioned and processed. Enter the
sample handling system (SHS), which must eliminate unwanted components, such as contaminants, that affect an analyzer’s performance, while keeping the components that need to be measured intact. “The required combination is analyzer technology with an effective SHS, and a controller with an easy-to-use operator interface as well as a comprehensive interface to test the automation system,” says Hill.
To this end, the company has developed flow controllers and system controllers, and recently launched its Horiba One Platform, which enables the integration of a greater number of instruments, analyzers (between 5 and 30) and SHS modules, organized into multiple lines of analysis – for example, tailpipe, engine out, mid-bed or diluted exhaust streams.
emission01emission02:
A fully equipped Horiba emissions test cell including an ADS-7000 Automatic Driving System. When integrated with appropriate measuring equipment, a chassis dynamometer and the test cell computer, ADS-7000 allows automatic laboratory operation

Comments Hill: “The system controller has to be flexible to handle a multiplicity of configurations within the system, and that’s what we’ve managed to achieve. It will enable us to fulfill future OEM requirements, which will be for more components with a greater capacity for customization.”
Meanwhile at UK-based Cambustion, engineering manager Chris Nickolaus highlights the move toward normalization of drive cycles, as part of the world harmonized light-duty test procedures (WLTP), as one of the industry’s current challenges.
“I’m running a WLTP on a chassis dynamometer,” says Nickolaus. “It’s more transient than NEDC, so it’s more demanding in terms of getting the robot or human to drive it accurately. But also in terms of optimizing the vehicle and the engine calibration, it needs a lot more time spent on optimizing the transient operation because you encounter a wider range of engine speeds and load operating conditions.”
In addition, Cambustion recognizes an increase in the use of automated mapping – placing an engine on a dynamometer and using a test matrix, while a computer optimizes all considerations. “For example,” says Nickolaus, “we have instruments running in test cells for three shifts a day, and the engine is being run at different steady-state and transient conditions to map all the emissions. Then you can choose different operating conditions or strategies to optimize. Automated mapping enables you to get a lot of data in as short a time as possible.”
emission03
Horiba’s chassis test cell features a chassis roller dyno for the emissions testing of passenger cars and other light-duty vehicles

emission04
The flexible engine test systems in Horiba’s Titan series offer extensive test functionality

emission05
Catagen’s first Maxcat unit has been bought by Mahle Powertrain

Dr Dean Tomazic, executive VP and CTO at FEV, argues that the most effective emissions testing takes into equal consideration the engine, transmission, vehicle integration and chassis development. “You have to work all technologies together,” he says. “In simulation you need to go through lots of scenarios to identify the best configuration that will enable you to meet your targets. There’s a lot of upfront work to configure the overall system and then optimize each component. The rest is calibration and application work, and from an algorithm development perspective, with new software in the controller, we look at airpath models to minimize engine-out emissions, as well as different direct-injection strategies and warm-up strategies to optimize the catalyst operations.”
Eric Watel, engine expert at Critt M2A, agrees that mastering huge product diversification is a major challenge in powertrain development: “To improve fuel efficiency and minimize engine-out pollutant emissions, the turbocharger matching must be perfectly optimized for each application. It requires a consistent application of simulation models and, therefore, thorough testing of the turbocharger to feed those models.”
Critt M2A’s turbocharger testing facilities comprise four development gas stands. “We are improving our experimental setup to extend the
turbocharger characterization,” says Watel. “The challenge is to implement industrial measurements that fit with advanced simulation models, only used in the academic world at the moment.”
Like turbochargers, aftertreatments have become increasingly important to passenger-car engines in recent years. Pi-Innovo, a services company based in the UK and the USA, has developed a range of one-dimensional mathematical models of common aftertreatment system components, such as DOC and DPF, taking input from conventional sensors and creating virtual sensors to understand what goes on inside components, including catalysts and filters. The models enable accurate control by predicting parameters such as temperatures, pressure drop, soot load and emissions in real time.
emission06
Chris Nickolaus and director Bruce Campbell with a test vehicle on chassis rolls at Cambustion. The company is based in Cambridge, UK

“Estimating gas species, temperatures and pressures inside the middle of the cat is necessary, says CTO David Price. “You can easily put in the temperature probes, but trying to get an undistorted sample of gas out of the middle is almost impossible without disturbing the flow. We use temperature samples at various points inside, and immediately in front of and behind each element of the cat.”
With the stringent Euro 6c legislation scheduled for full implementation in passenger cars by September 2018, the future will bring greater demands to measure particle numbers and small size particles for particulate matter. Price foresees further challenges: “The particles are so small, the question is how to trap and count them. Plus, they have a habit of coalescing and then breaking down along the exhaust.”
He also predicts that their impact on the public’s health will come to the fore: “There will be studies that show they have even more effects on health than thought. But the health problems they cause aren’t necessarily related to what comes from the tailpipe; the problem is when they get into the air and react with other things.”
Another expected trend – perhaps still five years away – is onboard measuring of real driving emissions (RDE) through portable emissions monitoring systems (PEMS). Cambustion’s Chris Nickolaus believes this will happen because “there are concerns that diesel NOx might not have fallen as much in the real world as hoped for. With RDE, you do all the usual development and certification, and then you’re checked up on in the real world against certain limits – but nobody knows what those limits are yet.”
emission07
Thermal testing of turbochargers at Critt M2A. Like Horiba, Critt will be exhibiting at Automotive Testing Expo in Stuttgart

He continues, “Getting any piece of equipment that’s used to working on a test bench to work onboard a vehicle is difficult. When you run on a test bench, you’ve got all the electrical power you want, but onboard you don’t want to load the vehicle with electrical systems because it skews the results if you start drawing power from the alternator. That means running off batteries, which means traditional vacuum pumps that produce a lot of heat are out of the question because they’ll flatten batteries very quickly.”
Pi-Innovo has already trialed this kind of testing, fitting analyzers into the back of a pickup truck. Price says the test was useful because it was
possible to have a range of drivers employ random driving styles while monitoring emissions. The measurements weren’t as accurate as in labs, but Price adds, “It’s better than simulation, because drivers do all sorts of things that can’t be anticipated in simulation.”
And Horiba’s Les Hill concludes, “To cater for the future trend of RDE using PEMS, our system solutions need the ability to integrate and exchange data with other test cells, such as chassis, powertrain, and engine. The test cells themselves already have a much wider range of environmental control – from sub-zero to above 40°C, for instance – and these will increase in number in the future.”

New kid on the block
Catagen made quite a splash at last year’s Automotive Testing Expo North America in Novi when it unveiled Maxcat, its first 200g/sec
flow-rate test rig for catalyst aging. The company has now sold the first Maxcat to Mahle Powertrain but it is currently at Catagen’s base in Belfast, Northern Ireland, running correlation studies and contract testing.
“Aging catalysts is a headache for the OEMs and this technology opens up opportunities for them to reduce costs, but gives them greater
flexibility in their testing as well,” explains Professor Roy Douglas of Queen’s University of Belfast, one of Catagen’s co-founders and a former
Ford and GM engineer.
The appeal of Catagen’s equipment lies in the energy savings offered over traditional engine-based catalyst aging methods. According to another of the founders, Andrew Woods, the company’s technology – based on Woods’ own PhD research – is up to 90% more energy-efficient
than using an engine to do the same test, resulting in operating costs up to 85% lower than engine-testing a cat, and produces 98% less CO2
at source. Catalyst characterization testing, such as light-off tests, can also be performed on the same equipment.
Correlation studies to demonstrate the capabilities and savings that are possible are now being performed on the Maxcat, which is flexible enough to reproduce the engine exhaust gas composition, flow rates and temperatures of OEMs’ individual aging cycles. The studies will complement similar data from the smaller, lower flow-rate Labcat unit. On completion of the studies, the Maxcat will be delivered to Mahle in Farmington Hills, Michigan.

International variations
“At the base level of equipment, there is a great deal of commonality,” says Horiba’s Les Hill about the global emissions-testing picture. “But it changes higher up in the process, such as when you look at the type and number of test cells in use. Some OEMs like to do as much testing as possible before the real vehicle is used – at the engine or powertrain level. Other OEMs tend to do less upfront and more at the vehicle level on chassis dynamometers. In addition, the type of analytical and SH systems you supply can depend on alternative fuel initiatives in certain countries, such as the use of CNG in Brazil.”
FEV’s Dean Tomazic adds, “Regulatory requirements dictate in many cases the use of specific equipment. An example of different measurement methods are those used for heavy-duty gaseous emissions, which are governed by 40 CFR Part 1065 in the USA. These highaccuracy requirements exceed the levels that European and Asian measurement systems have to demonstrate.”
And Chris Nickolaus of Cambustion highlights a Russia-specific manufacturer test that demonstrates some practical differences: “Because
of low ambient temperatures, there’s a test conceived around a parked vehicle with a gentle tailwind. Customers idle their cars to provide cabin heat, often for prolonged periods. They want to check that the combination of the engine and the aftertreatment system are capable of tailpipe emissions that couldn’t cause CO poisoning, if the tailwind blows the exhaust gas toward the cabin air intake, for example.”
emission08
A natural gas test cell at FEV North America. FEV has seven engineering centers across four continents

4T60E TCC Torque Converter Clutch slipping

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The customer came in with the complaint of poor fuel economy and noticed a difference in the tachometer reading during a steady cruise. This problem was very intermittent and happened when the engine was warmed up after a long drive. After climbing a slight grade, the transmission would down shift to 3rd gear and disengage the TCC (Torque Converter Clutch). After the grade was passed the TCM (Transmission Control Module) would try to reengage the TCC. The check engine light would flash and set DTC PO740 (TCC Slip), the TCM disengaged overdrive and TCC engagement for the rest of the ignition cycle. The customer would notice at a 60 mph cruise the tachometer was almost 3000 rpm. When the transmission is working properly, the tachometer should read approximately 2500 rpm at 60 mph. We hooked up our diagnostic software for a test drive and tried to capture the data line. Using the software gave us a good look at what was happening. Figure #1 is a recording of the drive from 0 mph to a steady cruise and then back to a stop.
40fig1

This transmission uses two solenoids to control the TCC application, the TCC On/Off solenoid and the TCC PWM (TCC Duty Cycle) solenoid. These solenoids are applied simultaneously. The TCC PWM solenoid goes to 100% (Duty Cycle) when the TCM first commands it on and drops to 0% (Duty Cycle) when fully engaged. The TCC PWM solenoid is used to cushion the engagement. When the TCC PWM solenoid is at 100% it is exhausting TCC apply fluid. When the TCC PWM solenoid is at 0% the TCC is applied full pressure. You can see on Figure #1 that every time the TCM shuts off the TCC On/Off solenoid the TCC PWM solenoid will be stroked to keep the valve body bores clean of debris.

Normally when the TCC is engaged the TCC slip should be low (0-20 rpm). You can see that this recording shows the slip in more detail (Figure # 2A).
40fig2a

Notice that Figure #2B has numbers attached to it for explanation. At frame 610 you will see a marker going up and down the screen. This mark represents the point on the recording at which the data is being displayed. Lets take a look at the numbers on Figure #2B and identify them.
1. TCC on/off solenoid engages
2. TCC PWM (TCC Slip) solenoid turns on. (100% = exhaust 0% = close)
3. TCC Slip is about 165 rpm. (Slipping)
4. TCC Slip is about 83 rpm. (Slipping)
40fig2b
Here's what we found at frame 640, the TCC remains engaged but the slip speed goes up. The throttle has been opened and engine load has increased at that time. We have used my scan tool in bi-directional mode to actuate the solenoids. They have a distinctive click when they are actuated, so I believe that electrically the circuits are ok.

It sounds as though you have a TCC PWM bore that's worn out. We frequently get this call on the HelpLine. The graphs you have displayed show the TCC on/off and the TCC PWM doing their job as far as the TCM is concerned.The interestingpart is how you determined the valve body bore might be worn. Knowing the solenoids are being commanded is half the adventure. Verifying the solenoid is working is the key. The bore being worn is virtually impossible to confirm outside the vehicle

We installed a new sleeve and converter regulator valve to the valve body bore. We captured another graphic after the repair to verify the fix. Figure #3 shows normal apply and application of the TCC PWM and the TCC on/off solenoids. The slip of the TCC RPM in conjunction to the TCC PWM is now on time and working correctly.
40fig3

Rocker Geometry 03

[falloutboy]

Der wohl beste Artikel den ich jemals über dieses Thema gelesen habe. Teil03

Installation, Measurement and Adjustment Tips
figure01
Here’s the typical stud mounted needle bearing roller tip rocker shown in its CLOSED position. Dimensions are not shown, because they are relative to your installation, which is relative to your NET valve lift (after lash, if used). The important dimension you want to find
and set, to be HALF your net valve lift, is the ROCKER HEIGHT, shown to the upper right, and illustrating the cumulative value between the ROLLER axis and TRUNNION axis. The other references are shown with regard to their heights above and below (respectively) the valve spring RETAINER. All measurements for where the trunnion is sitting, is referenced to the top of the retainer, marked here by TRUNNION REFERENCE. Raising and lowering your rocker’s tail by adjusting the Pushrod as needed, will set this. Although shown in this illustration, the adjuster doesn’t
even need to be here, as it will only get in the way. Let the rocker sit loosely on the study, with its adjustable pushrod and set this trunnion reference as needed to get the trunnion exactly HALF of your NET valve LIFT.
If using a HYDRAULIC TAPPET, be sure that it is fully extended during this check. You can prime the motor to do this. After getting your pushrod length, ADD .020” to allow for hydraulic tappet compression during actual engine operation. Order EXACTLY what you need for pushrod length, rounding off to the nearest ten-hundredth of an inch (two decimals).

figure02
This is the references to pay attention to for setting a SHOE tip rocker arm for your NET valve lift. As mentioned, finding the centerline of the fulcrum’s rotation is required first, since this cutaway only shows where the axis would be. This is not easily seen on the outside of a stamped or cast rocker body so, you need to simulate this rotation and make a mark to reference to.

figure03
As crude as this may look, this is an actual recreation of the drawing illustrated to author by the late Harland Sharp, explaining his original layout of a rocker silhouette on paper before scribing where the roller would be, and using an actual roller on his outline to calculate
this. Setting the roller’s diameter in direct position where the scuffing surface was at, instead of the roller’s centerline is what is wrong here. It was the beginning of a duplicated error that would last over a quarter century.

figure04
As explained in text, this is the real example of what you don’t want to do. But it is very typical, and the problem is propagated more by the head companies luring engine builders into a false sense of “acceptability” to such an installation, by selling their heads with studs in them that
have this stretched over placement, when they’ve had to move the pushrod for wider ports. The better alternative would be to leave the STUDS on the same centerline as the VALVES, and force stud rocker manufacturers to put their offset for the pushrod solely in the rocker arm,
with an offset cup. This is what Ford did on their N-Head, and it is the best way. Otherwise, you need a stand (shaft) system. Here’s the bottom line: You can never have an inline valve array cylinder head, like a SB Chevy, Ford, Mopar, etc., and NOT have the rotating axis of the
rocker’s trunnion be IN LINE with the CAM. Any twist at all, is COMPOUND geometry, and will make it impossible for the roller to lay flat on the valve, or follow the correct straight down path on the Y-axis.

figure05
This represents the STAND (SHAFT) rocker system, with only the most important things to be considered; namely, the STAND, the SHAFT and the ROLLER TIP.
The valve is shown for angle and location, and is shown here in the MID-LIFT point of motion, as noted to the side. THIS IS THE GOAL. The
roller and shaft are to be horizontally in line with each other as measured to a perpendicular right angle with the valve.

figure06
This shows a SQUARE (crosshatched) being used to lie against the valve and atop the shaft of your stand system. The stand is bolted to the head, and a shaft is laid in position to now make this check while the valve is in the CLOSED position. The cool thing about this is it can be done on a work bench. No springs, no anything, just the parts shown. If you are doing this with an assembled head, you can run the square
along the side of the valve springs, providing they’re uniform diameter. Otherwise, you may have to use the valve spring retainer technique from our stud mount example.

figure07
Here’s the stand (shaft) mount system shown in the close position, and our example here has a NET valve lift of .650”, or a MID-LIFT value of .325”. You MUST know this for your engine. It is impossible to set correct mid-lift rocker geometry without knowing your net valve lift. The ROLLERS are shown here in their two critical states, the dashed version representing where the roller will be when it has opened the
valve to FULL lift. But the valve is only shown in the closed position, as is the solid roller atop it.
First, take HALF of your ROLLER diameter; and HALF of your SHAFT diameter, and ADD them together, you will come up with a “standard.”
In this example, that standard would be .521”. This comes from a roller diameter of .480” and a shaft diameter of .562”. Why half? Because this finds our CENTERLINE for each. It is always the centerline that you are setting with rocker geometry.
Second, write down the height of the SHAFT’S TOP above the valve tip.
Third, you need to write down your MID-LIFT value (half net valve lift).

Here’s the TRICK:
With these three things written down:

1. Subtract HALF net valve lift from your STANDARD (.325” from .521” in this case, for a sum of .196”).
2. Subtract the sum of the above (.196” in our example) from your rocker’s height (for our example this would be .350” minus .196” = .154”. This is how much the SHAFT needs to be lowered to bring the centerline of the shaft, half of the net valve lift below the center of the roller). Usually, with shimming, you might have figures that make you ADD shims to get the correct value.

As you might notice, we are using the outside diameters of the shaft and roller to calculate this standard from, because these are easily measured with common tools. But it is their centerlines that are so important.

Jim Miller has been involved in the racing industry for more than 30 years. Jim’s roots stem from amateur racing, starting at 16 but by age 17, he was the youngest Ford trained line mechanic authorized to do warranty work on all of Ford’s factory muscle cars. At the age of 21 Jim was invited to take over Chief Mechanic duties for Dyno Don Nicholson’s Pro Stock Maverick, which Jim respectfully would decline – passing the opportunity on to Jon Kaase. Jim holds a number of patents for valve train design and is the proprietor of MID-LIFT Precision Geometry. Jim can be reached at 1777 Blount Rd., 501 Pompano Beach, FL 33069. Phone [phone].

Rocker Geometry 02

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Der wohl beste Artikel den ich jemals über dieses Thema gelesen habe. Teil02

Setting Geometry
As mentioned, the rocker geometry has TWO considerations: The accuracy of how it is installed, which I have always referred to as the “installed geometry.” And second, how accurately it is designed, which I have always referred to as the “design geometry.” This second item relates to where the adjusting screw or cup is placed in the body, and at what angle.
As an engine builder, you can move the rocker arm up and down to the valve tip in setting that side of the rocker’s effect, but you can’t do too much about the pushrod side, and that is where the information comes from. That is up to the rocker arm manufacturer.
Of the two sides, however, the lesser of evils to shoot for is setting the VALVE side (or installed geometry) as closely as possible, because (a) this is where the motion is constrained by the valve guide, (b) this is the side that has the greatest motion, (c) this is the side that has the valve spring, where harmonics are generated and amplified, (d) and this is the side where all of the foregoing multiply into a measurable resistance value that generates heat, robs power and creates additional friction. The pushrod is, by comparison, free floating with the rocker body’s movement in and out as well as up and down, and it is moving less (the cam’s lift). But make no mistake about it, when the pushrod side is not performing to mid-lift geometry, it is losing information. The upside though, is that whatever is left, is getting through to the valve. When the geometry is off on BOTH sides of the rocker, because you didn’t install it accurately, you lose twice! Only 90% may go into the rocker from the cam, and of that, only 90% comes out, or 81% goes to the valve. I’m of course rounding things off for example, but the principle is the same.
I should add one other point here. Everything is “net.” So if you have those cute little “checker springs” laying around, find some other use for them, because outside of holding a valve in a head for display where someone can use their finger pressure to push the valve open, they have little use. You need the REAL running springs for any geometry setup. The same goes for checking flex or piston to valve clearance or anything else critical. Checking springs ADD about .040” (or more) NET valve lift to your engine. Or, another way to say it is you will LOSE .040” or more NET valve lift when you put the heads together with the running springs, compared to whatever you measured using the checking springs. This is true across the board, flat tappet cams, roller cams, aluminum rockers, steel rockers – it makes no appreciable difference.
Later in this article, I will offer installation and assembly tips but for now, here is the easiest of accurate ways to set INSTALLED rocker geometry:
The closed valve position is the easiest and the best. The cam must be in the closed position, on its base circle. Heads are assembled to the engine, with no pushrods in place. You must know what your “net” valve lift is supposed to be (we can get nitty-gritty later). You will subtract any valve lash so you have an accurate “net” lift.
For stud rockers, put it in place with an adjustable pushrod. You don’t need the poly locks; just let it set loosely on the stud. Knowing your net valve lift, DIVIDE it in half, and write it down on a piece of paper. Then, lengthen or shorten the adjustable pushrod to raise and lower the back of the rocker until you get the center of the trunnion exactly HALF of your net valve lift BELOW the center of the ROLLER TIP. If for example, you have .600” net valve lift, then this would be .300”. Keep in mind that I refer to “center” of the trunnion and roller pin. It is their axis that is what you are measuring. Some are easy to see and some are not. For those with flat machined surfaces, take a scribe, measure and mark these centers as best you can. But the main trick is that you want to be sure you measure this from a precise 90 degree reference to the valve centerline. To accomplish this, you are best served to use the top of the valve spring retainer. Simply lay a short machinist ruler (or something comparable) atop the valve spring retainer, and pass it along the side of the rocker arm.
When you have the height installed accurately, the trunnion will be exactly HALF of your net valve lift, below the roller tip centerline when the valve is CLOSED. As it opens, and moves to exactly mid-lift, the axis of the roller tip will have dropped down to be straight across from the trunnion and an imaginary line that runs between them (I call the motion line) will form a precise right angle with the valve centerline. The
roller is at its farthest point across the valve when this happens. When the valve continues to open the second half of its lift, to full lift, the roller will have moved exactly an equal amount BELOW the trunnion as it was above the trunnion when it was closed. And the roller will
be at its closest inside point on top of the valve. You will also have the ultimate least amount of roll across the valve.
For shaft mount rockers, it’s a little different, but the principle is the same. With shaft rockers you must use shims, or have a stand that has a surplus of metal that you are machining exactly what you need away. But you can take a measurement of the stand height without using a rocker arm. Just bolt the stand down to the head with a couple of bolts, lay the shaft in the stand’s bed-way, and use a machinist square along the side of the valve (or spring) and shoot the long end along the top of the shaft, so there is a gap beneath it and the top of the valve. Everything is about finding the centerlines, and being creative about doing this, while being accurate at the same time and measuring at an accurate right angle with the valve.
It makes no difference where the wear pattern is at on the top of the valve, when you have correct mid-lift geometry, and providing the pattern is “on top of the valve.” (Running off the edge of the valve is not good.)

Graphing the Cam at the Valve
To see what is really happening at the valve, you need to check your rocker’s motion by measuring it at the valve. The best way is to essentially treat this like you’re degreeing your cam, but you’re measuring motion at the valve. Only instead of just picking up points of lift
to compare to the crank, as you would with the .050” tappet lift measurements on a cam card, you will be creating a graph all the way through the entire cycle of valve lift, opening and closing. If you are fortunate to have a CAM PRO or CAM DOCTOR, or something similar, life is good. If not, you can do the old fashioned way. You will need graph paper that can be found at art stores, engineering supplies and many science or school supply providers. You need a dial indicator that you’ll mount directly above the valve spring retainer, nearly fully compressed so you follow the valve’s stroke fully – and you want to be sure the indicator stem is lined up parallel with the valve. As with setting your cam, you need to have a degree wheel in place on the crank and zeroed accurately to TDC of the piston.
With the above in place and ready, you have TWO choices to how you measure this; which are merely opposite perspectives. You can choose an even number of crank degrees you move through to measure valve lift, or you can choose an even number of valve lift to measure degrees of crank rotation. It doesn’t really matter which you use, because it is the comparison against other like tests that is important, and
both need to be the same. You don’t have to be too crazy about fine increments here, just choose valve lift jumps of maybe .020” and write down the crank movement; or choose 5 or 10 degree crank measurements and write down the valve lift.
Personally, I like the second method of using a fixed crank stepping, and then noting the valve lift. This goes directly to the points I make about “area-underthe-curve” that you are looking for. For those new to this term, area-underthe-curve refers to the VALVE LIFT CURVE as charted across a graph of time (meaning crank degrees), and what most engine builders agree, is that lifting the valve as quickly as possible and as soon as possible, while setting it down quickly after it has hung open for as broad a period of time as needed, but without being too fast to damage the valve train from excessive “bounce” is what everyone wants. So, when you want to see the gains and losses in this area from inaccurate rocker geometry, you’re really looking for wasted time when the crank is moving more than it needs to lift the valve the same amount. So, if you standardize your testing to the same valve lift measurements, the gains and losses in the crank are readily seen.
Once you’ve charted one rocker geometry setup, perhaps the one you’ve been using, now make the changes with pushrod length and/or stand height (for shaft systems) to meet with what I hope I have informed about earlier in this article. You will often see that PEAK valve lift is very close to the same, but much less at other points in the curve. That is the lost information. The degree of this will depend on many factors that take another story to itemize. But the bottom line is; you will appreciate how important it is keeping the same geometry, for making sure your cam changes show results that are directly accredited to them and only them. Otherwise, your information is tainted.

Shoe Tip Rocker Geometry
As with aluminum rocker arms, there are different design geometry shoe tip rocker arms, but the priority for adjusting the valve tip side is still there for the same reasons. However, you don’t easily have the same accuracy as you do with picking up precise center points on the roller and trunnion of a needle bearing roller tip rocker arm. Even finding the axis is not easy. Setting it by the same rules is best simply for standardizing one cam to another. Only your reference is the actual contact point of the pad itself. When it is at mid-lift, you will have a 90 degree relationship between the TIP of the valve and the center of rotation for the rocker arm. But finding that center is tricky, because these are usually ball fulcrum rockers, and they are surrounded by the stamped metal that has no clear axis to it. One solution is to put a stud upside down in a vice and rotate it carefully while observing the fixed point on its side that most closely represents where the center is, then making a little felt tip pen mark. This would then be set exactly half of your net valve lift below the valve tip in the closed position. It’s not as accurate
as fixed points to set calipers against, but it will get you very close if you have patience and a sharp eye; and with shoe tip rockers, the amount of error you might be off will have no measurable effect in cam efficiency as it would with needle bearing roller tip rockers.

Twisted Rockers
Unfortunately, engine builders are led into a false security by stud mount rockers sold for heads that shouldn’t use them. These are aftermarket heads with pushrod offsets that require an offset pushrod cup, or adjusting screw. Shaft systems usually have this adequately
fixed, but when heads are sold with studs in them, that clearly need to be removed for a shaft system, this is the false sense of security one gets. The two rockers shown here is exactly what you DON’T want to have (Figure 4).
When you have this much rotation, the roller tip does NOT lay flat on the valve tip, and as it opens the valve its energy is making a cross sword slash across the valve tip that rounds off the top of the valve tips, side loads the guide on the X axis (length of head), and shifts side loads to the bearings in a way that often push or break one prematurely. It’s bad news, costs horsepower and breaks parts.

Ratio & Geometry
If the rocker geometry is off, then so too is the ratio. There’s some good news though: Don’t worry about it, because very few of the rocker arm makers did. The history of rocker design didn’t have much accuracy involved. There was no standard, because there was no need for such accuracy in the old days. Rockers always were, and to a great degree, still are designed in the closed valve position. But as long as you stay focused on what you need to do, you won’t try juggling things around outside of the real priorities, based on a false idea of what YOUR rocker ratio ought to be. It can be all over the place from .05 less, up to maybe close to what it is supposed to be, simply because many manufacturers began in the closed position, and then started moving specs around from a hit and miss until it got close. Once it did, they left it there. Over the years, more consistent manufacturers with the least amount of broken parts have been the model other newer companies would copy. But the mistakes get copied, too. You have to always check things for yourself, forget the advertising. If you do, then you can’t blame anyone but yourself.

Rocker Geometry 01

[falloutboy]

Der wohl beste Artikel den ich jemals über dieses Thema gelesen habe. Teil01

Rocker Geometry
by JIM MILLER
rocker01

ROCKER ARM GEOMETRY seems to raise its head every now and then, and when it does, I rarely ever see it stated accurately. Too often a sound bite of only a small piece of information is taken out of context and then used as the Gospel, totally ignoring the other dynamics that revolve around it. In some cases, something totally erroneous is stated that is not only wrong, but makes no sense for anyone who just stops and thinks about it objectively.
When lecturing at trade shows, schools, engine shops or just getting pinned down on the phone by a knowledgeable engine builder going deeper than most on a technical issue, I have found that I spend about half my time trying to undo various misconceptions about rocker geometry
before I ever begin explaining the facts. There has been so much info put out there by reputable companies (and by my reckoning, incorrect), that people are reluctant, by nature, to see something different from the prejudice of what they already know or think they know. If people are used to doing something a certain way, they see everything from that perspective. Usually, my getting through to them involves discrediting what I think is wrong with what they were doing and then begin to explain what they needed to change. At that point I could break down the simple rules for what geometry really is, and why.

Background of Rocker Geometry
Rocker geometry (or the lack of it), goes way back to many fathers on both sides of the ocean, to when the Wright Brothers were still studying the theories of lift in an airfoil. But for our purposes here, and to avoid boring the curious who’ve managed to get this far on this story, I’ll come to the point about rocker arms and explain as needed how the “mistakes” got to where they did.
In the old days, rocker arms were all pretty much what we term a “shoe” design; meaning the contact pad with the valve had a large radius scruff surface that depressed upon the valve tip as the rocker moved through its rotation. The term of course comes from the appearance to a shoe’s sole, but also to the mechanical motion much like a foot and boot would do, as it pushes off. This pushing off motion, as many will already know, has the effect of the rocker arm stretching itself as it moves through the depressing (lift) cycle. It is actually lengthening itself as it moves across the valve tip, and you see this by the wide foot print (we call a “witness mark”) atop the valve tip.
The use of rocker arms goes back to many things predating engines, but the principles were never required to be so specific on axis point heights and their consequences, as it is for helicopter bell cranks, and racing engines! There was no rocket science to designing these parts a hundred years ago, which ended up on our prehistoric cars and early airplanes. Engineers simply made designs that tried to minimize the degree of how much scuffing was imposed on the valve tip; got it close, and moved on to more important questions. Somewhere along the line, there became a principle to get this in a general ballpark, that someone later coined as the 1/3-2/3 theory (or either of the two). This placed the pivot point of the rocker arm so that it was 2/3 of the way below the valve tip, or the valve tip was 1/3 of the way above the
rocker shaft, depending on your point of view. But the answer was the same. This thinking was originally derived from the intention that a near 90 degree arc could be realized when the valve reached its intended full open position. Bear in mind that valve lifts back then were usually
in the quarter inch or so range, on little two and four cylinder engines. So being off a little really had no measurable difference in performance of the engine, and wear and tear was the real yardstick engineering back at the turn of the century was aimed at. Also, the ability to accurately measure wear and tear, horsepower, thermal loss and many other cool things we take for granted on today’s computers, wasn’t even a possibility back then.
The advent of more valve lift, and thus pushing a budding internal combustion engine technology higher to produce more power was really inspired for leaps and bounds by the advent of aviation, not Henry Ford. Not to take anything away from the automotive crowd’s contributions, but only aviation imposed the second requirement that defined “efficiency” – and that was light weight. Making Goliath engines that had more power was a lot easier than making more power from light weight engines that would be flying over somebody’s head, somewhere. So the
whole “thinking process” for efficiency in engine technology really found its impetus in aviation, because racing back in the early 1900s was still done on the back (or behind) of one-horsepower whose exhaust was more easily stepped in than emitted from a pipe. And as far as my 30 year old memory serves me on the research, aviation was also the first use of a “roller tip” rocker arm, on radial engines as far back as the 1930s, and perhaps before. In fact, to this day, I never cease to be amazed at the foresight and creativity of both aviation and automotive engineers in the 1920s, and ‘30s, and ‘40s. Four valve Pent roof combustion chambers, roller cams, fuel injection, nitrous oxide, water injection, two stage superchargers, turbo chargers, and many other cool things we assume were concepts of the last 20 or 30 years, were actually done and done quite well, seventy and eighty years ago.
Aside from the roller tip rockers of aviation long ago, the fundamental rules of rocker arm design were based around the shoe tip, contact pad design still used today. Many of you may know that you can’t use a roller tappet on a flat tappet cam, and of course vice verse not only because of hardness of material difference, but because of “geometry.” The principles of trailing motion and dynamics between something that is making direct LINEAR contact upon another object that is imposing or receiving a RADIAL (circular) is entirely different than if that contact is occurring with a roller tip making or following the contact. This isn’t rocket science either, and you can see how this happens by drawing a roller tappet in various stages of lift as the cam lobe goes around to push upon it, and see a straight line from its axis to the cam lobe is constantly
shifting around as the tappet goes from the close (base circle) position, up along the acceleration ramps, then over the nose. When finally, as it crosses dead center at full LOBE lift, this straight line between its axis and the cam centerline is also in alignment with the tappet bore itself. At all other times, the tappet is actually receiving some level of side thrust in its bore (engine block) from the pushing out forces that the lobe imposes as it chases it up.
Even when engineers were chasing efficiency with aircraft engine and power development, the threshold for seeing measurable loss of engine life, like valve stem or valve guide wear, was not easy unless things were really out of whack. They weren’t worried about loss of cam
events through small changes in rocker geometry, even though they knew the variables existed. So if they kept to this 1/3-2/3 rule, everything looked good on the valve tip, and the leverage of the rocker arm upon the spring – or more accurately stated – the leverage of the
spring on the rocker arm was at near perpendicular relationship with the valve when rates were at their highest. Even though by today’s standards for valve springs, spring rates back in the 1920s and ‘30s were negligible. This “attitude” continued on throughout the decades
afterwards. More importantly, and unfortunately, it bled over into the soon to come roller rocker arm market, that got its impetus in the 1950s. The first person who made a working aluminum “roller tip” rocker arm for automotive application belongs to my dear old and
departed friend, the late Harland. Some other garage efforts might have been getting tinkered with out in California about the same time, but it is pretty well undisputed that 1958 is the beginning of what we know today as “the aluminum roller rocker.” Keep in mind that aviation
roller rockers existed twenty or more years before, but they were steel, they were radial engines, and they didn’t comport to the automotive needle bearing aluminum body that Harland introduced.
Just like the flat tappet cam and the roller tappet having entirely different geometry because of where the measurement for motion is made, the same rules apply to the shoe tip rocker versus the roller tip rocker arm. But when Harland made his silhouette, he didn’t allow for this, and inadvertently moved the axis of the roller roughly .300” of an inch higher than what it should have been. The axis of the roller should have
been in the same place as the contact pad. So when his rocker was placed on the engine, and the roller was positioned for a good “eye-ball” track on the valve, now the push-rod cup was too high. The result, was that it went way up and in toward the stud. In actuality though,
most engine builders in the sixties continued to keep using standard length pushrods, and the excessive motion from this mistake was occurring on the valve tip, which was deemed “normal” because the roller rolled! Believe it or not, even to this day, people think that the roller
tip is for rolling on the valve. It is NOT. The roller tip is for one reason only, and that is to convert the shifting length of the rocker’s arc (that moves across the valve on a shoe design), to a fixed length that moves far less in its effect, because it is always point down in line with the
valve’s motion, just as a roller tappet of a cam is always aiming its contact tangent line with the axis of the camshaft.
This error stayed, and was copied by many manufacturers and eventually by everyone in some measure or another. It would take several decades before enough trial and error, and even a patent would be studied to make manufacturers rethink this, slowly improving their
designs. Ironically, some of the most well known names continue to promote designs they never changed, and even promote the less accurate means of using little tools, that tell an engine builder what pushrod they need, while never even taking into account the valve lift that will be used. Make no mistake; you cannot set rocker geometry without knowing exactly how much the rocker arm is going to move.
It seems logical that since a roller cam can provide all the acceleration any of today’s heads need, for any rocker geometry scenario, then why not set the rocker geometry to ONE STANDARD that has the least amount of wasted motion, and will always duplicate the same percentage of cam information, regardless of what cam you use? For understanding this, understanding a little history is always best. This ends a lot of rhetoric.
Whenever pushrods leave their inline paths to now have their end follow around the rocker shaft by any amount, this is LOST CAM INFORMATION. The cam literally has to turn more degrees to affect the same LIFT at a later point on the crank. Velocity, too, is lost. So you’ve lost duration, throughout the entire lift cycle (not just overall), and you’ve lost RATE of acceleration, by slowing the rocker down.
To put this in perspective, let’s take a simple even value of cam lift, like .400” to make a point. Rocker geometry is usually thought about as only what is happening at the valve. In our .400” cam lift example, and a 1.50:1 ratio rocker arm, this would theoretically yield .600”. To have the optimum use of in-line motion being converted into circular motion, you need to divide these values into two equal parts. Engine builders do this at the valve, but the real deal is happening with the CAM. The cam is the source of the information. So for the CAM side of this value, we’re only ending up with .200”, which is one half of our .400” cam lift. If you fail to place the axis of the tip of the pushrod at the proper length as to divide that cam lift accurately, so the rocker arm is at a 90 degree angle, then you have a pushrod that is going to move in and out more than it needs. The result will be wasted cam information that can require the crank to move several degrees more to effect the same lift of the valve. Those lost degrees were absorbed in the excessive motion the rocker arm had. You will spend hours and hundreds of
dollars to get a camshaft that is ground to fourth decimal accuracy, and chosen to give you a specific degrees of duration at .050” tappet lift, and you will change a cam to gain as little as four or five degrees if you think the engine needs it, but you just threw away more than that
because of the PIVOT POINT on your rocker arm didn’t establish the correct angles with the pushrod and valve.

Why did this continue?
Back in the early 1960s, because there was so much inefficiency and experimenting with cylinder heads, cams, induction systems and so on, this valve train flaw went under the radar. Now I was still a punk, barley sixteen in 1969, but if memory serves me right, it wasn’t until a real student of engineering and racing stuck his nose into the situation, and started shifting the pivot point around for his own purposes and seeing
distinct changes. He had some odd name car in a new class of drag racing, called Pro Stock; I think it was “Grumpy’s Toy.” Bill Jenkins was one of the real pioneers for many things in not just fixes to problems, but also a more scientific Yunick-like approach to analyzing. He didn’t follow other people; other people went out of their way to follow him. It was a short list of real pioneers to both cylinder head and valve train
development back then, and Bill was on the short-short list. But prior to Bill and a few others like him of that era, rocker geometry was totally ignored beyond the vague generalities of the 1/3 rule. But technology in the cams and heads was soon catching up. Right about this time,
in 1969/70, Chrysler approached Crane Cams for a new camshaft for the factory backed Hemi teams of Sox & Martin, Herb McCandless and the “Motown Missile” (later Mopar Missile). That development was the beginning of the .700”-plus valve lift boundary being broken.
The late sixties and early seventies were really exciting times for factory muscle cars, and the stepping stones of technology that has brought us to where we are today. It all began back in this limited, golden era. And the fundamentals established then, cut in stone, have not changed to this day either. They’ve only gotten repackaged, renamed and resold, even though other boundaries in valve lift, cylinder heads and so forth have been elevated. The principles for cam technology and specifically rocker arm geometry that would soon come along in 1980, but spawned in 1973, have not changed to this day.

Definition
What is rocker geometry? Rocker geometry is “angles of motion.” It is not some linear reference point on the tip of the valve, that trying to adjust the wear pattern will guarantee it being correct. What is correct? Correct, is “efficiency.” It is having the least amount of wasted motion being used to do the greatest amount of work (that is designed to be done by the cam). This last point is important, because the rocker arm
can be used to add to the cam, besides what it usually does by error, which is take away from the cam. But I will get into each of these below. I just wanted a simple “mission statement” that defines what geometry is and is not, so that the following hopefully makes sense.

Importance
WHY is rocker geometry so important? When you change the pivot point of where the rocker arm is, in relation to the valve tip, it CHANGES THE CAM. It doesn’t matter whose rockers you use, it doesn’t matter what style rocker you have, it doesn’t even matter what application your engine is; whenever you change location of the rocker pivot point in relation to the tip of the valve, you are changing cams. You are changing all three parameters simultaneously: LIFT, DURATION and VELOCITY (rate of acceleration).
The degree you change these depends on how much you move the pivot point. And one or two of these three parameters may be affected more than the other. But if you don’t LOCK your geometry in to the SAME THING all the time, which has the least amount of wasted motion, then you are aiming at a moving target with every cam change. Whatever results you get from one cam to another is tainted by the diluted effects of wasted motion in the rocker arm.
Rocker arms are a “radial” device being ordered to do a “linear” thing. They rotate on an axis, moving in a circle. But what they have to impart is a straight line command. They get their order from the camshaft, in the form of IN-LINE information that they then have to ROTATE around an axis and then MULTIPLY it by some ratio, and finally TRANSFER this result back to another IN-LINE component of greater movement. This movement has THREE values: LIFT, DURATION (of lift), and VELOCITY (acceleration of lift). If the rocker arm does ANYTHING ELSE
besides this, then it is NOT efficient, and SOME of this information is being lost.
Let me make a point about something on this. Your camshaft is ground to ten-thousandths of an inch precision. It is computer designed to millionths of an inch, you (or your cam manufacturer) selected it for a division of duration values where you considered two or three degrees important; any more was too big, and any less was too small. Hopefully by now, you realize that moving this rocker pivot point will change this at the valve.
You just don’t know how much. Well, it is MORE than the two or three degrees you think is important. In some cases it can exceed TEN degrees, and is often five or six degrees. As if this isn’t enough reason, consider this: It is that value of loss throughout most of the lift cycle, not just total – where you’re only inclined to measure it from, and where your cam card is limited to. That is what’s missing. Engine builders only check at FULL LIFT. They check rocker ratio and total valve lift, and that’s that. But when your rocker geometry is off, you’ve lost those
degrees of duration throughout most of the entire cam profile. Which means rate of acceleration is lost, but you may only see a small change in lost valve lift, thinking the difference is just flex in the rocker ratio.

Two Geometries
Rocker Geometry is the correct DESIGN and INSTALLATION of the rocker arm so that its relationship to both sides of up-and-down motion is fully realized by BOTH. This is of course, the pushrod, and the valve (respectively).
The rocker arm is a RADIAL tool, asked to do a LINEAR job. It pivots around an axis in a reciprocating radial (circular) motion, and has all the dynamics that anything revolving around an axis will have. But on either end of the rocker arm’s connecting points are two other instruments each live and breathe by the laws of linear (in-line) motion.
Now all this may sound like “rocker arms 101” and we may all know this, but few people I have found over 35 years, seem to understand how sensitive, precise and important this observation is. I think this true because most treat both the design of the rocker arm and its installation with casual regard.
To have the “most efficient” design and use of the rocker arm requires TWO things: (1) The rocker arm must be designed to mirror the inherent angles of each engine’s pushrod to valve geometry. (2) Rocker geometry must have an accurate location of its rotational axis with the valve tip’s height. The first of these is called DESIGN geometry; and the second is called INSTALLED geometry.
Every engine has an inherent acute angle (I call the “attack angle”) which the pushrod leans either into or away from the valve centerline. The small block Chevy for instance is 19 degrees positive (leans into). This comes from the engine block having a 4 degree angle of its tappet bores with the piston cylinder centerline. The head is already a 23 degree valve angle (actually it is a 67 degree, because it references from
the deck), so you simply subtract the 4 degrees that the tappet and pushrod “aim” toward an already inclined 23 degree valve, and you end up with simple math: 19 degrees. Every engine is unique. The SB Ford is 20 degrees for this same value. This is what rocker geometry needs to be designed to, otherwise your efforts of installing the rocker arm accurately will be limited to just one side of the rocker or the other.

What Rocker Geometry IS NOT
Rocker geometry is NOT where the roller (or contact point) is at on the top of the valve. Forget that. Rocker geometry, especially is NOT the idea that you want to place the roller or wear pattern (shoe tip rocker) in the “middle of the valve.” Forget that, too.
The valve tip is everything. It is ground zero. This is where all leverage, and the full stroke of valve lift begins, this is our reference point. There are many ways to measure rocker geometry, but there is ONLY ONE way to SET IT. Now, you can set rocker geometry in the closed position, or you can set it in the MID-LIFT position (half open), or you can do like most people have done, and simply roll the engine over a couple dozen times watching the valve open and close to see your witness mark (foot print) atop the valve tip and play hit and miss with chasing a “minimal wear pattern.”
The problem with this latter point is that this is a symptom, it is not geometry. Granted, when you have the rocker geometry set properly, you will have the least amount of wear pattern, but to try to set geometry through moving the rocker and trying to see how small you can get it, is better than nothing – not quite good enough. You can easily be off by .005” to .010” (or a great deal more) on even seeing this actual width, let alone measuring it. And being off .010” on the horizontal plane of what you are really trying to measure, which is the vertical
plane (valve lift), will multiply out quite easily to .030” or .050” or .080” or more in your error of where the trunnion is to the valve tip! Those kinds of errors will cost you several degrees of crank rotation to open the valve a like amount.
What about using a tool, or dial indicator designed to measure this in and out motion, resting on the spring retainer? Well, this is a better way of the same thing, but it is still measuring a horizontal plane for a vertical plane result. Error can be off several times more than measuring directly in the vertical plane, or parallel to the valve motion itself. When using these tools, just like a dial indicator on the top of a piston, as soon as the piston reaches perfect TDC, you will have two or three degrees of crank movement before you see the dial indicator move. There is a float time there, and so too is the effect by using a tool on the roller tip of a rocker to measure in and out motion. It floats enough to allow the valve lift to be off by .005” to .010” or more. But if you can set it dead nuts within .002” to .003” without having to buy such a tool, why wouldn’t you do it in a more precise way?

Alternative Geometries
Before getting too deep into philosophies, history and facts, let me restate the key point of what rocker geometry is, then I will mention the comparative arguments people (and companies) have made against this. MID-LIFT geometry is rocker geometry that has the ultimate “efficiency,” in that it is doing the greatest amount of work with the least amount of effort. It has the least amount of wasted motion in the pushrod and valve, commonly referred to as the “inand-out” motion. It affects the maximum response through linkage of whatever the
cam’s instructions are. If you don’t have geometry set precisely, your consequences range from simply losing a little cam information at the valve, to excessive side loads in various directions, on various parts that will at the very least rob you of power. In more extreme cases, wearing out parts or outright catastrophic part failure may occur.
There are arguments to using different geometry than mid-lift. I simply don’t agree with them because they violate the principles of efficiency. One of these theories is to adjust the rocker arm’s height so it reaches a 90 degree relationship when the valve is about 2/3 open, not half open. The logic being, that the spring loads on the rocker body are less. Another reason I’ve heard is that it accelerates the valve in the “mid-range” better, thus making more power. Other variations of this approach shift the rocker arm’s pivot point higher on the valve tip to create this 90 degree effect sooner in accelerating the opening of the valve at a lower point of lift, thus increasing what is termed “area-underthe-
curve.”
Both of these are a way to add different cam information to the valve by using the rocker arm’s geometry. The only reason you would use the rocker arm for creating a “second dynamic” of valve acceleration, is if the cam was unable to give you the acceleration you needed. Now, in some cases this limitation exists. They would be flat tappet cams, of mechanical or hydraulic operation, and an engine where cylinder heads had the flow potential, and/or cubic inches had the demand which required a crazy acceleration to midlift flow values off the seat. In other words, the engine was so big, and the heads were so big but for rules or some other illogical reason, the cam they HAD TO USE was a flat tappet profile that had limited “rate of acceleration” by its limited tappet diameter and base circle constraints. I won’t get into cam technology and limits, but that is the first reason I can think of for using the rocker arm as its own cam tuner. In some Stock classes where the original cam must be used, a creative (and well funded) engine builder can play games with rocker geometry to change valve lift rates, but
these are very limited differences, usually not worth the trouble, and most of all, in ALL these examples, there is going to be detriments that outweigh benefits.
In the first place, for those who have an engine of large flowing heads, and big cubic inches, or heads for very high rpm’s, they will have the benefit of using a roller cam. So the issue of how fast you can open the valve is not even a consideration, because by nature of roller tappet geometry, any value of acceleration up to and through suicidal parts destruction can be implemented on the cam profile. And for those classes
where a flat tappet cam is required, the cubic inches and head limitations of most rules I’ve seen over the years, fall within airflow and rpm limits that a flat tappet cam fit just fine. Too many times, cam companies talk customers into roller profiles that are not needed, and in fact don’t make as much power as a well chosen flat tappet would, because it takes more power to operate the roller. Using rocker geometry as a second cam shaft is not a good idea. The velocity of the rocker arm increases where its motion line reaches a 90 degree angle, and trying
to pick a particular segment of the valve lift that you want to impose that thinking over what the cam manufacturer has done, is bad news. But there’s another point to consider on this issue.
The rocker arm is a symmetrical device to whatever geometry it is set at. In other words, whatever acceleration it exhibits on the opening side of the valve gets reversed on the closing side. Simply, if you set geometry with a HIGH pivot point, so it increases its velocity quickly off the valve seat, then slows to full lift... Guess what? It’s going to accelerate back to the close position when it leaves full lift. Because it will always mirror whatever its settings are.
To see the difference all you have to do is take an old fashioned needle pointer torque wrench and turn the engine over without spark plugs and measure the drag. Then set the geometry to MID-LIFT, and see the difference. I’ve had people do this and fall out of their seats. They see as much as 45 foot pounds or more and LESS torque to turn the engine over. Usually it is 15 or 20 foot pounds but it depends on how much spring pressure you have, how complex the rocker geometry is (Hemi versus SB Chevy, for instance), and how wrong their geometry really was. Either way, it quickly shows you the definition of efficiency. If they MAP their valve acceleration, throughout the entire lift curve, then I need say no more. The best way is to DIVIDE the arcs EQUALLY and standardize this for all cam and cylinder head testing. Change your cam
as you need.

Schrauben - Typen, Arten und Nutzen

[falloutboy]

Zitat:

@spechti schrieb am 11. Februar 2015 um 17:10:41 Uhr:

Einleitung
Jedes amerikanische Auto bis in die frühen 80er hinein hat zöllige Schrauben. Komplett und ohne Ausnahmen (naja, bis vielleicht auf die 14mm-Zündkerzengewinde). Danach wurde schrittweise und bei Einzelaggregaten auf metrische Verbinder umgestellt. Allerdings ist die Umstellung bis heute nicht vollständig abgeschlossen, da das SAE-System in Nordamerika stark verankert und überall präsent ist. Der Grund für das zöllige Maß ist die Herkunft aus der angelsächsischen Maßwelt – obwohl gerade diese kurioserweise heute längst größtenteils metrisch mißt. So sind England oder auch Australien längst metrisch. Kanada ist ein Mischfall. Eigentlich wird metrisch gemessen, aber durch die Nähe zu den USA sickern viele Einflüsse des SAE-Systems durch tägliche Gegebenheiten wie den Autoverkehr ein. Speziell in den USA (und in einigen wenigen kleineren Ländern) hat sich das metrische System bis heute nicht durchgesetzt und ist lediglich durch Importgüter wie Autos oder Haushaltsgeräte präsent.

Auch bei uns ist das metrische System in Bezug auf Schrauben nicht omnipräsent. Z.B. durch die amerikanisch geprägte Computertechnik kommt zölliges Schraubengut zu uns. Zöllig ist nicht gleich zöllig, da gibt es noch Varianten. Ich werde mich hier aber auf die im US-Autobereich gebräuchliche Schrauben beschränken. Auch britische Gewindevarianten wie das Whitworth-Gewinde werde ich weglassen.

Grundlagen
Was im deutschen das Maschinenschraubengewinde ist, ist im amerikanischen das Unified Standard Thread. Beide Systeme gibt es als Grob- und Feingewinde. Bei der Zollvariante redet man dann von UNC (Unified National Coarse = grob) oder von UNF (Unified National Fine = fein).

Vergleicht man die Maße der metrischen und zölligen Gewinde miteinander, wird man keine Übereinstimmung feststellen. Durchmesser, Steigung und Gangzahl sind anders. Die einzige Gemeinsamkeit ist bei beiden der Gewindeflankenwinkel von 60°. Die US-Gewinde sind aber nicht kompatibel zu beispielsweise den britischen Whitworth-Gewinden, die trotz - oberflächlich betrachtet – gleichem, zölligen Gewindedurchmesser aber einen Flankenwinkel von 55° haben und somit völlig inkompatibel zu Unified-Gewinden sind.

An amerikanischen Autos sind fast nur Schrauben mit Grobgewinde (UNC) oder grobem Blechschraubengewinde im Einsatz. Feingewinde kommen auch vor, sind aber deutlich in der Minderzahl. Sie sind beispielsweise an Vergasern zu finden.

Einer der Vorteile des zölligen Systems ist die bei Ami-Cars sehr begrenzte Zahl von verwendeten Schrauben-/Gewindegrößen. Man kann so ein altes US-Auto mit einer sehr übersichtlichen Anzahl von Schlüsselgrößen nahezu komplett zerlegen.

Trotz genormter UNC/UNF-Gewindeformate wird man bei genauer Betrachtung feststellen, daß sich 99% der an US-Autos verwendeten Schrauben grundlegend von gewöhnlicher Baumarkt- oder Schaubenhandelsware unterscheiden. Diese Unterschiede sind spezielle Zugeständnisse an die einfache und schnelle Handhabe am Fließband oder an einen besonderen Zweck (wie zum Beispiel Festigkeit oder Einbau an einem unzugänglichen Ort). Auch ist die Qualität von richtigen Autoschrauben um einiges höher als die von Schüttware.

Systematik
Üblicherweise wird der Durchmesser eines Gewindes mit einem Bruch dargestellt, z.B. 1/2 oder 3/8. Bei Schrauben werden so gut wie nie Dezimalzahlen verwendet, also nicht z.B. 0.5 inch.
Alle Schraubengrößen sind Fraktale von 64, wobei 64 der größte verwendete Nenner ist.
Kürzen der Brüche ist üblich. Somit wird aus 16/64 dann 1/4 inch.

Schrauben unterhalb von 1/4 werden nicht in Brüchen dargestellt. Dafür verwendet man einfach natürliche Zahlen von 0-12. Eine Schraube #12 ist also eine Größe unterhalb von 1/4. Eine #0 ist sehr klein. Die Kennzeichnung, daß es sich um eine Zollschraube handelt wird mit obenstehenden Anführungszeichen deutlich gemacht, z.B. 1/2".

Man sieht also: man sollte in Mathe bei Bruchrechnung nicht gepennt haben! Ich werde Rechnen hier weitgehend vermeiden, aber man sollte halbwegs wissen, was ein Bruch ist.

Nomenklatur
Um eine Schraube genau zu bezeichnen sind mehr Angaben nötig als nur der Gewindedurchmesser. Daher werden bei einer Schraubenbestellung auch noch andere Daten genannt.

Eine davon ist die Anzahl der Gewindegänge pro Zoll. Diese wird dem Gewindemaß mit einem Bindestrich nachgestellt. 3/8-16 beispielsweise bedeutet, daß ,man eine Schraube mit 3/8 inch-Gewinde und 16 Gewindegängen pro Zoll vor sich hat.

Eine weitere angefügte Information ist die Schraubenlänge, die ebenfalls in Zoll angegeben wird. Diese kann entweder ebenfalls al Bruch dargestellt sein, oder aber auch als Dezimalzahl. Beispiel: 3/8-16 x 1 ½ oder 3/8-16 x 1.5

Die Benennung von Feingewinden ist von der Sache her das gleiche. Eine 3/8-Feingewindeschraube fällt in der Benennung sofort durch ihre höhere Gangzahl auf, die bei 3/8 lautet: 3/8-24 x 1.5

Bei allgemeinen Zollschrauben ist das bereits alles an notwendiger Information.

Da es – wie oben bereits erwähnt – für Autos besondere Schrauben gibt, sind bei Autoschrauben noch besondere, weiterführende Informationen in der Bemaßung hinzugefügt. Das kann z.B. ein Teilgewinde sein, eine aufgepreßte Unterlegscheibe, ein spezielles Material oder eine Oberflächenvergütung. 3/8-16 x 1.5 wsh 1.0 bezeichnet eine Schraube mit 3/8-Gewinde, 16 Gängen pro Zoll, anderthalb Zoll Länge und einer Unterlegscheibe von einem Zoll Durchmesser.
Auch Informationen zum Schraubenkopf können vorkommen. „hex“ z.B. bedeutet einen Sechskantkopf.

Da bei komplexen Schrauben an Autos die Benennung ellenlang werden würde (zu lang für viele Manuals) haben Ford und GM eigene Nummernsysteme für Schrauben eingeführt. Bei GM ist es eine der Teilenummernlogik folgende 10-stellige Nummer, bei Ford ist es ein Buchstabe, Bindestrich und eine fünfstellige Zahl, z.B. S-54735. Wird in einem Manual irgendwo eine Verschraubung dargestellt, steht diese Nummer daneben. Zu Chrysler habe ich keine Informationen.

Beispiele
Eine originale Ford-Karosserieschraube hat meistens 5/16"-Gewinde, es gibt auch welche in 1/4". Bis auf das Maschinengewinde hat sie nichts mit einer gewöhnlichen Maschinenschraube gemein. Sie hat eine spezielle Führungsnase (dog point) um an unzugänglichen Stellen die Schraube besser einführen zu können. Zur besseren Druckverteilung ist eine Unterlegscheibe daran. Diese ist vor dem Rollen der Gewinde aufgesteckt worden und kann nicht abgezogen werden. Der Grund dafür ist, daß Scheibe nicht unbeabsichtigt abfallen soll.
Auch der Sechskantkopf hat eine Form, die es nicht unbedingt erfordert, den Schlüssel gerade aufzusetzen um die Schraube zu drehen. Diese Schrauben wurden entwickelt um Autos am Fließband schnell bauen zu können.
ford-karo

Sie sind normalerweise aus hochwertigem Stahl gemacht und am Ende der Produktion phosphatiert und in schwerem Maschinenöl geölt. Sie sind zwar nicht wirklich rostfrei, lassen sich aber im Regelfall auch nach 30 Jahren noch problemlos rausdrehen. Nach einem gründlichen Reinigen kann man sie oft wiederverwenden, so robust sind sie. Hergestellt wurden und werden sie von der Firma SEMS. Zu erkennen ist das Herstellerzeichen oben auf dem Sechskant.

GM hat ähnliche Schrauben für Karosserien entwickelt. Ihnen fehlt allerdings die Führungsnase. Stattdessen hat man die Schrauben vorne spitz konstruiert und das Gewinde bis auf die Spitze gezogen. So hat die Schraube auch bei nicht ganz richtigem Ansetzen bereits Grip und die Form hilft auch in gewissen Grenzen gegen das verkanten.
gm-karo

Sonderformen von zölligen Gewinden:
NPT
An amerikanischen Autos befindet sich noch eine weitere zöllige Gewindeform, ohne die nichts läuft und die ich kurz beleuchten möchte: das zöllige Rohrgewinde NPT (National Pipe Thread).
npt

Genau genommen ist es eigentlich keine Schraubenform, sondern eine Methode, Leitungen und Anschlüsse dicht miteinander zu verbinden. Dies sollte in einem Produktionsprozess so zügig und sicher funktionieren, daß kein zusätzlicher, zeitraubender Dichtmittelauftrag erforderlich ist. Erreicht wird dies durch eine kegelige Gewindeform, die sich keilförmig ineinanderdreht und fest dichtet. Dieses Merkmal unterscheidet das NPT-Gewinde von den Schraubengewinden.

Ein weiterer Unterschied ist, daß ziemlich offensichtlich ein 1/4“ NPT-Gewinde in der Größe stark von einer 1/4"-Maschinenschraube abweicht. Das NPT ist bei gleichem Nennmaß wesentlich größer. Der Grund dafür ist historisch bedingt und stammt daher, daß ursprünglich eine andere Betrachtung des Rohrmaßes zugrunde gelegt wurde.

NPT-Gewinde kommen am Motor vor, aber auch bei Kühlsystem, Getriebe und Kraftstoffanlage. An Autos ist nur eine kleine Anzahl an NPT-Gewindemaßen üblich. Häufig sind 1/4“, 3/8“ und gelegentlich auch 1/2“.
NPT-Gewindetabelle
Es gibt auch im britischen Whitworth-System ein kegeliges Rohrgewinde mit den Whitworth-üblichen 55° Flankenwinkel, welches aber in Autos keine Anwendung findet (vielleicht bei Engländern. Kann ich aber nix zu sagen). Dies ist nicht kompatibel zu NPT.

SAE
Oft verwechselt wird das NPT-Gewinde mit dem SAE-Gewinde, das beispielsweise bei Bremsleitungen verwendet wird. Dieses hat im Gegensatz zu NPT kein Kegelgewinde. Männchen und Weibchen lassen sich gerade ineinanderschrauben. Da SAE-Gewinde bei Rohrverbindungen nicht dichtend sind, wird die Dichtfunktion über einen Bördel erreicht und das Gewinde lediglich für den Anpreßdruck genutzt. Bei SAE-Gewinden wird – wie bei Maschinenschrauben – das Zollmaß und die Anzahl der Gänge pro Zoll angegeben. Der Bördel und der Dichtkonus weisen einen Winkel von 45° auf.
Am gebräuchlichsten an PKWs sind die Größen 3/8-24, 1/2-20 und 7/16-24. Diese werden für Bremsleitungen mit 3/16“ Durchmesser (in Deutschland schrägerweise 4,75mm genannt) benutzt.
sae

AN
Eine weitere – relativ neue - zöllige Rohrverbinderform, die an Autos vorkommt, ist der AN-Verbinder. Das Buchstabenkürzel steht für Army/Navy und bezeichnet ein ursprünglich militärisches Anschlußverbindungssystem, das nach dem zweiten Weltkrieg seinen Weg über die Rennszene ins Auto gefunden hat. Es basiert auf dem SAE-System, ist aber dazu wegen einiger Unterschiede nicht kompatibel. So weisen z.B. Konusse von Bördelverbindungen einen anderen Winkel auf. Dieser beträgt im Gegensatz zum o.g. SAE-Bördel nur 37°. Durch den spitzeren Winkel entsteht eine größere Dichtfläche und ein höherer Anpressdruck. Eine irrtümliche Verbindung zu einem SAE-Bördel würde eine leckende Verbindung ergeben. Daher sollte man das strikt trennen und nur die für Übergänge vorgesehenen Adapter verwenden.
an-elbow
Das AN-System hat seine eigenen Benennungen, z.B. wird eine Verbindung mit AN-8 oder AN-6 bezeichnet. Das System ist in Schritte von 1/16“ aufgeteilt. Somit ist z.B. ein AN-8 = 8/16“ gekürzt also = 1/2“.

Möchte man AN-Leitungen in sein Auto einbauen, sind Adapter nötig um z.B eine Kraftstoffleitung an eine Benzinpumpe anzuschließen. Die kompletten Programme der Firmen Russell oder Earl’s bestehen aus AN-Teilen. Sie sind im Custom-Bereich allgegenwärtig. Es gibt auch immer mehr Benzinpumpen, Vergaser, etc. von Holley, Edelbrock und Co., die bereits AN-Anschlüsse haben. Ihr Vorteil ist ihre strikte Systematik (man braucht nur einen AN-6-Schlauch und einen AN-6-Verbinder zu kaufen und hat Gewißheit, daß diese Teile perfekt zusammenpassen), ihre universelle Anwendbarkeit und ihre hohe Qualität (aus dem Luftfahrtbereich)
AN-Gewindetabelle

Schrauben messen
Nachdem wir nun einige Schraubenarten und einige Grundformate kennengelernt haben stellt scih natürlich die Frage: wie finde ich heraus, was für eine Schraube ich vor mir habe?

Zu diesem Zweck eignet sich am besten eine Schieblehre. Es gibt diese in digitaler Form oder mit einer Skala. Ich habe zwar schon digitale gesehen, die auch Fractions in Zoll anzeigen können, die sind aber recht selten. Daher zeige ich Euch die einfacherer Variante mit einer Skalenlehre.

In diesem Fall habe ich eine Lehre mit zwei Skalen. Eine metrieche in cm/mm und eine in inch.
schiebskala01
Auf dem Bild ist die Zollskala oben im Bild zu sehen.
Bei genauer Betrachtung ist zu sehen, daß die Skaleneinteilung nicht auf dem Dezimalsystem beruht (es also 10 Striche bis zur "1" sind), sondern 16. Aus Sicht des unbedarften Europäers taucht sofort die Frage auf: WAS SOLL DAS???

Bei etwas drüber-nachdenken ergibt sich daraus, daß die Skala somit bis zur 1"-Marke 16/16 hat.
Demzufolge ist der längere Strich in der Mitte zwiachen der "0" und der "1" gleich 8/16", gekürzt also 1/2".
Die Schlußfolgerung daraus ist, daß die Striche auf halbem Wege zwiachen der "0" und der 1/2"-Marke dann die Viertel und Achtel sind

schiebskala02
Wir spannen eine Schraube in die Meßbacken der Lehre. Die Meßmarke bleibt über dem fünften Sechzehntelstrich stehen.
Die Schraube hat also 5/16"

schiebskala03
Eine andere Schraube wird eingespannt. Hier bleibt der Strich auf 8/16" stehen.
Hier wird der Bruch gekürzt und ergibt somit 3/8".

TIPP: Es ist schwieriger, das ganze dezimal zu messen und dann die 6,74mm in einen Bruch umzurechnen. Daher ist es praktikabler, auf Umrechnungen ganz zu verzichten und komplett in der zölligen Welt zu bleiben.

Vergleich Karosserieschrauben
Um noch mal deutlich den Unterschied zwischen GM-Schrauben und Ford-Schrauben zu zeigen, habe ich mal aus der Schraubenkiste drei Schrauben gefischt, die verschiedenes deutlich zeigen.
schrauben01
Links:
Ford 5/16" Karosserieschraube aus meinem Ford Mustang, hergestellt 1970.
Sie hat 40 Jahre unten am Auto gesessen, dort wo der Kotflügel an den Schweller geschraubt ist.
Sie hat eine aufgepreßte Unterlegscheibe und den "dog point", die Zielhilfe.
Die Schraube ist nur grob mit einer Plastikbürste gereinigt und es ist deutlich zu sehen, daß die Phosphatierung noch vollkommen intakt ist. Sie ließ sich ohne Probleme lösen und könnte sofort wieder geölt und verschraubt werden.

Mitte:
Eine GM-Karosserieschraube in 5/16" von 1977. Innenkotflügel an Außenkotflügel Buick Century. Deutlich erkennbar ist der Kopfflansch und der gespitzte Gewindeteil.
Es ist deutlich zu sehen, daß die Schraube ein- oder mehrmals mit metrischem Werkzeug verrundet wurde. Trotz der unvorteilhaften Behandlung ist auch diese Schraube problemlos zu lösen gewesen (allerdings schon vor etlichen Jahren) und auch hier ist die Schraube durchaus nach Reinigung und neuem Ölen im Prinzip noch verwendbar.

Rechts:
eine metrische Schraube M6 aus einer Corvette, mitte der 90er hergestellt, 2001 aus dem Auto entfernt. Frontmaske. Schon nach wenigen Jahren starke Korrosion am Kopf und an der Unterlegscheibe, Gewinde gerade noch akteptabel und drehbar. Reinigung ist nicht mehr Möglich, nur durch Strahlen. Diese Schraube würde ich nicht wiederverwenden.

Man sieht deutlich, wie sich die Materialien mit dem Niedergang der Fertigungsqalität in den 80ern und 90ern massiv verschlechterten. Ich habe neulich an einem 2000er PT Cruiser geschraubt. Die Schrauben sahen noch schlimmer aus als die der Corvette.

Vergleich U-Nuts / J-Nuts
Im Karosseriebereich von US-Autos gibt es kaum Sechskantmuttern. Es werden fast überall J- und U-Nuts eingesetzt, gelegentlich auch einfache Blechschrauben. Der deutsche Begriff lautet Klammermutter. Der unterschied zwischen J- und U-Nuts ist die Länge der Schenkel. Bei J-Nuts sind sie unterschiedlich, bei U-Nuts gleich. Es gibt auch noch weitere Varianten davon.

Wenn mal eine Schraube wirklich festgammelt, sind oft nicht die Schrauben schuld, sondern die U-Nuts. Auch bei diesen geht die Qualität weit auseinander. Auch dazu habe ich ein Bild.
j-clip
Links:
Sehr große, neue U-Nut, chinesische oder mexikanische Fertigung. Größe 7/16". Für Stoßstangenbefestigung.
Bei genauem Hinsehen ist sie nur brüniert oder KTL-grundiert (kathodische Tauchlackierung). Obwohl sie neu ist und noch nie verbaut war rostet sie schon.

Zweite von links:
CIP-Nut, neu , 5/16" für langen Blechübergriff aus aktueller Ferigung. Etwas besser, aber auch eher lasche Qualität, trotz Markenhersteller. Kann man verbauen, ist aber im Außenbereich am Auto nicht wintertauglich.

Dritte von links:
CIP-Nut, NOS aus Fertigung USA 70er Jahre in 5/16" mit kurzem Blechübergriff. Kleine Lagerungsspuren und etwas staubig aber solide und intakte Phosphatierung. Würde ich sofort verbauen. Hält geölt locker noch 50 Jahre.

Rechts:
CIP-Nut neu in 1/4" mit kurzem Übergriff. Qulität ähnlich wie die lange 5/16".

Nachtrag:
Noch diese PDF zu AN-Fittings auf der Festplatte gefunden (siehe Anhang ganz am Ende des Beitrags)

Nachtrag 2:
Über diese Bilder gestolpert, die hoffentlich ein paar der gebräuchlichen englishen Begriffe bei Schrauben, Muttern und Unterlegscheiben erklären.
headstylebolttypesnuttypeswashertypes

Friction elements - Product Profile

[falloutboy]

BORGWARNER’S LATEST FRICTION PLATES ARE IMPROVING AUTOMATIC TRANSMISSION EFFICIENCY

In the development of new vehicle components, manufacturers are obliged to consider customers’ needs, such as fuel efficiency and shifting
comfort. Current trends in automatic transmission design take these considerations into account and focus on reduced package size, increased
torque density and durability as well as improved efficiency and shift quality. Suitable friction enablers with great reliability, even at high temperatures, are essential for this.
Today, modern 8- to 10-speed transmissions require friction materials that can handle increased power density and higher energy levels. In conjunction with the necessity for minimal dimensional changes in the lining, smoother clutch engagement demands a consistent positive ?-v relationship of the friction torque curve throughout the entire life of the transmission under various operational conditions. The friction performance of a wet-clutch system is affected by the chemical and physical interactions of various fluids with friction materials. Wet friction
elements are used in shifting or starting clutches and have to provide stable friction characteristics as well as high temperature resistance.
They must be able to handle limited oil flow and higher torque, which can be generated using higher unit loads or materials with high friction
coefficients. In addition, simulation tools help to reduce drag torque by optimizing the friction plate design.
Being based on specific requirements, the design of the friction plate is unique for each application. On receipt of the Statement of Requirements (SOR) from the customer, the first steps in the design process comprise the verification of the geometric layout by calculating the net pressure
on the basis of the required torque capacity and the thermal calculation of the interface and oil outlet temperature according to the shift cycle specified by the customer. After drag torque calculation using analytical, CFD and neural network simulation tools, the fourth step covers a lifetime prediction based on duty-cycle data and durability testing at different energy levels. All steps result in the definition of the friction plate design with regard to lining/friction material, groove geometry, core plate geometry and manufacturing in terms of segmenting and post processing.
Modern friction materials provide improved heat resistance even at lower cooling flows, thus facilitating safer operation over the entire lifespan. Also, they necessitate further and continuous advance, such as friction elements with a high surface-adsorption capacity. These readily adsorb the oil friction modifiers in the automatic transmission fluid (ATF) while not being affected by degraded ATFs.
BorgWarner’s newly developed family of friction materials helps to fulfill these requirements. Specifically designed for wet starting clutches,
torque converter lock-up clutches, torque transfer clutches and hybrid disconnecting clutches, the BW 6910 friction material provides
resistance to oil degradation and glazing, withstands high interface temperatures and maintains a stable and positive ?-v characteristic.
It thus enables a low-lube concept to be used, and the reduced cooling oil flow permits the use of more efficient pump systems and optimizes
transmission efficiency. In addition, the friction material facilitates the handling of higher surface pressures and a reduction in the number
of surfaces or the friction diameter. Vehicle measurements of shudder at acceleration under micro slip conditions verified the advantages
of this material in contrast to standard launch friction material.
Recent automatic transmission architectures demand a higher differential speed on the shifting clutch elements. Extreme shifting conditions at 70m/s and an extremely low oil flow can cause a severe accumulation of hot spots on separators. Specifically for shifting clutches in modern automatic transmissions, BorgWarner developed the new BW 5000 friction material family, which is extremely elastic, has a uniform oil retention surface and features a high-temperature fibrous surface.
Reducing drag losses inside a wet clutch plays an important role. In a dual-clutch transmission, for instance, drag losses can be classified into three sections: losses occurring during pre-selection, idle-D losses and drag losses at the seal rings or bearings. Nevertheless, most of these
negative effects can be reduced or avoided by optimizing the software of the transmission control unit or via implementation of an engine stop/start system. In addition, an optimized clutch design as well as improvements in the applied friction material are additional possibilities
for reducing drag torque.
Various calculation tools can be used to predict drag torque reliably. Analytical modeling, for example, uses a calculation program with adjustment based on actual measurements. The neural network method applies artificial intelligence that depends on previously collected
data, while a third method using CFD software performs exact calculations of fluid behavior inside the clutch. The latest calculations resulted in
the following friction material design solutions for effectively reducing total drag torque: waved friction plates, waved separator plates, an
optimized groove pattern, active separation and two-step lining.
The core plate design offers further potential for optimization. One new concept is the ‘hemmed spline’ design, in which a doublefolded core plate steel in the spline area increases the spline contact area, thus allowing the clutch pack length to be reduced without reducing the contact area of the splines. In general, hemmed spline design enables a reduction in weight, axial space and material costs.
A further modification method is the segmentation of the core plates. This facilitates the more efficient use of the core plate steel and is
therefore a reasonable cost reduction effort. BorgWarner started series production with an OEM customer in 2012.
Recent transmission design trends necessitate improvements in friction materials. Choosing the appropriate friction product depends on the
individual application and must be predicted using simulation tools in close cooperation with transmission and vehicle manufacturers. These
simulations help to predict the drag torque and enable optimization of the friction plate layers during the design phase.
BorgWarner’s BW 6910 friction material allows clutch systems with a low-lube strategy, provides a high torque density and improves
durability as well as NVH robustness. Also, advanced friction materials, like the BW 5000 family, allow shifting at high differential speeds and prevent the severe accumulation of hot spots on separators. Drag torque, packaging and costs can be further improved by the use of new design concepts, such as the hemmed spline design and segmented core plates.
fig01

Platzierung/Dimensionierung Motorkühler

[falloutboy]

Cooler by design
Formula 1 teams have had to go on something of a journey of discovery when it comes to cooling. Cranfield University student Pierre Salmon looks at how computational fluid dynamics can help the teams to extract the most from their cooling systems

Cranfield University’s Advanced Motorsport Engineering MSc allows students to get closer to Formula 1 through interaction with professionals in the industry, detailed academic courses, a group design project and an individual thesis. One of the 2014 theses looked at evaluating the required air mass flow rate through the sidepods needed to reject the waste heat produced by the F1 powertrains to comply with the new 2014 regulations. Computation fluid dynamics was used to investigate the potential cooling and aerodynamic benefits of five different
configurations and any effects they might have on the performance of the engine.

The 2014 F1 technical regulations concerning powertrains have provided considerable challenges to the engineers regarding the packaging of the cooling system. The addition of a turbocharger has resulted in more required heat rejection to the air flowing through the sidepods. The oil and water cooling requirement for the engine remains relatively the same for the 2014 V6 engine compared with the 2013 V8 engine, but the addition of the charge air cooler results in a much higher requirement of cooling air mass flow. There is also an increase in electronics
cooling requirement from the higher power outputs of the MGU-K and more complex energy management electronics. A strong need therefore exists to find a solution that could possibly reduce the cooling requirements or find a cooling configuration that allows for the highest heat rejection rates.

Designing the most efficient cooler configuration has ample benefits as it affects the three performance differentiators of the 2014 season – power, aerodynamics and reliability. Firstly, effective cooling of the charged air reduces the density of the air going into the cylinder. This allows more molecules of oxygen per unit volume to be reacted and means more fuel can be combusted per cycle, allowing for a higher IMEP. Secondly, the e ective cooling of the air, water and oil will reduce the average operating temperature of the engine and so extend the life of the engine. The reliability and life of all the components in the engine is crucial to the successful operation and racing of the car since only  ve powertrains were allowed per season per driver in 2014.

The enlarged sidepods of the 2014 cars provide a major reduction in aerodynamic efficiency of the vehicle as they slow down more air and reduce the clean  ow of air to the rear of the vehicle. It would therefore be of benefit to any team to be able to increase the heat transfer abilities of the cooling systems and reduce the cooling air mass flow rate.
figure1
Transferring heat
An increased efficiency of the engine fluid cooling will also free up cooling capacity for the electronics and turbocharger. Throughout the 2014 pre-season testing it was found that efficient cooling of the cars’ systems resulted in major gains in reliability and performance, and the first half of the 2014 season saw numerous iterations of bodywork to optimise the cooling of the vehicles. In this project, the cooling from
an air to air charge air cooler (CAC), a water radiator (WRAD) and an oil radiator (ORAD) are considered. The typical mass flow rates of air
through the sidepods would be between 1.2 and 2.1kg/s, depending on vehicle velocity and sidepod inlet area. The 2014 engine needs to reject around 113kW of heat from the water radiator, around 41kW from the CAC and in the region of 58kW from the oil radiator.
The five different configurations see the CAC, WRAD and ORAD placed in a way as to maximise the heat transfer or reduce the external pressure drop across them. Extreme Temperature Difference (ETD) is the main driver behind heat rejection from coolers and designers typically ‘sweat off’ the heat as much as they can by staggering the hotter surfaced heat exchangers behind the colder ones as in
Configuration 4. Another strategy is to have all heat exchangers exposed to the cool inlet air and size them relative to their heat rejection requirement, as per Configuration 2.
graph1
The geometry of each heat exchanger is optimised to provide similar performance while maintaining a specified volume. It is often the greatest challenge for aerothermodynamicists to package these relatively bulky devices around the engine and in the sidepod. Dense
cooling cores are used to maximise the surface area and fin efficiency, with fin density being as high as 21 fins per inch. The fin efficiency
of each core is also crucial as this allows for greater heat transfer to the cooling air. It is also important to understand how the heat exchangers affect the performance of other heat exchangers downstream. If you imagine the heat exchangers are packed very close together inside the sidepod, then once the air has passed through the front heat exchanger, it has warmed up and is traveling slower and is more laminar, all of which negatively affects heat transfer. Traditionally louvres are used to break down the aerodynamic and thermal boundary layers along the  fins to re-establish a high temperature gradient and turbulent flow near the heated surface.
graph2
Front wheel tyre wake is also an issue as this highly turbulent air, which is moving at a slower relative velocity, will inhibit heat rejection performance of the cooling system in the sidepod if it were to enter it directly. Aerodynamicists now cleverly use the Y-250 vortex to help divert the wheel wake around the lower side of the sidepod. Not only does this help with reducing drag, but it also allows for faster cleaner moving air to enter the sidepod and effectively, and predictably, extract heat from the cores. See Figure 1.

The CFD process used a generic sidepod geometry generated in Catia V5, with the five different cooling cores input as finite tubes of
varying temperature. There were in excess of 25 million cells in the meshes and the simulations took four days each to run on the Cranfield
High Performance Computing Cluster. The macroscopic heat exchanger model was not used because obtaining the heat rejection to mass flow rate correlation curves from an F1 team are like trying to find hen’s teeth. In addition, the separation events and effects of tube geometry was interesting to see, so the full detailed CFD was done in order to gain as much information as possible.

What is interesting to note is that Configuration 4 yields the highest heat rejection rate at the car’s maximum velocity of 90m/s. This is due to it having the cooler CAC in front of the warmer radiator, so the absorption of heat into the cooling air is maximised. The pressure drop across the configurations are also of interest because it is no good having a cooling system that rejects the heat but costs a significant internal aerodynamics drag penalty. Now, what is of further importance is to observe what the air is doing after it has left the cooling cores and how this affects the rear of the vehicle. The air velocity uniformity at the sidepod outlet is of concern because it affects the aerodynamics at the rear of the car. The sidepod inlet shape determines the air mass flow distribution and thus the outlet flow condition. The ideal situation would be to have uniform air mass flow distributions through the cooling cores, and no severe pressure gradients at the outlet. Observing the rear of the sidepod (so looking forward towards the outlet) we can see a comparison of the outlet velocity streamlines of each configuration.

Rotational flow
The flow structure behind Configuration 3 might appear more chaotic, but of importance is the strength of the rotational flow present
behind all the cores. Configuration 1 shows a double counter-rotational structure, with the primary stronger vortex on the right-hand side, which develops initially due to the non-uniform pressure distribution aft the cores. This induces the left-hand side counter-clockwise vortex due to viscous sheer forces between particle layers.
figure2
Configurations 4 and 5 show similar trends in a strong rotational  ow structure aft the cores, where the region of low mass flow rate results in lower particles per volume, which directly relates to a lower pressure, which accelerate high pressure particles region towards it. The high pressure particles already have momentum towards the sidepod exit, and their gradual acceleration towards the low pressure region results in the rotational flow structure being established.
figure3
Observing the Q-Criterion 0.018, vortex laments show how the rotational flow structures develop in the sidepod after the cores. Configuration 3 shows smaller and fewer vortices than Configuration 1. The counter-rotating vortices of Configuration 1 are shown, and the two larger vortices labelled near the top of Configuration 3 dissipate due to longitudinal acceleration as the cross sectional area decreases near the outlet.

The vortex exiting on the left (1) is produced from, initially, the flow over the last row of tubes at the bottom of the water radiator
rolling over and creating a standing vortex. As the air  ow from higher layers exits the cores it follows the  ow and encourages the rotational
 ow laterally (around the axis indicated by the dashed arrow), which wraps around and then rotates around an axis perpendicular to the page as the vortex nears the exit, due to the lower pressure aft the core from the non-uniform pressure distribution mentioned before. The difference in strength between the vortices of Configuration 4 and 5 is that with Configuration 4, the flow encounters the longer radiator tubes last, and so the flow is disturbed less than with Configuration 5 with the shorter CAC tubes at the rear. The less disturbed flow of Configuration 4 has a higher velocity (boundary layer build up and effective flow area) and the higher the exit velocity the higher the tangential velocity of the vortex. The higher tangential velocity, due to Newton’s first law, results in a larger radius, but this also increases the volume of the vortex and a lower core pressure is experienced, increasing centripetal acceleration, Newton’s Second Law. The equilibrium state between the two forces results in a faster spinning stronger vortex, observed by Configuration 4.
figure4
Hotspots
The right-hand side vortex is also rotating clockwise, but is slightly weaker, and in the case of Configuration 5, is dissipated before the outlet because the flow accelerates in the longitudinal direction as the cross-sectional area decreases. The core regions of these vortices (inside the  laments) are relatively calm, where viscous dissipative forces damp out any large scale turbulent fluctuations. Generally this helps preserve the core region of the vortex and helps extend the life of the vortex. The resulting pressure contours at the outlet is shown below, where Configuration 3 has the lower pressure gradients due to more uniform flow, and Configuration 4 and 5 have the strong vortex’s low pressure in the filament core. As a result of the sidepod outlet flow conditions, it could be speculated that in the future teams might want to make use of this flow in their aero packages.
The thermal energy exchange between the heated surfaces and the cooling air can be gauged by the temperature differences of the air after the cores. The following images show the temperature rise inside the sidepod. Configuration 3 clearly shows a hotter region just behind the oil radiator as the air is being heated by three stacks of cores. A noticeable difference between Configuration 4 and the other configurations is the uniformity of the air temperature after passing through the cores.
In contrast, Configuration 1 and 3 have regions of higher temperature and regions of lower temperature, showing variation of the
mass  ow rates through the cores. Ideally, equal mass flow rates of air through all the rows of the cores should be achieved with optimisation of the sidepod inlet shape, plan view shape and exit duct shape. This image clearly shows the temperature gradients and streamline flow into
and out of the cooling cores.
figure5
The affect of these CFD results were then applied to an engine simulation of a 2014 regulation F1 engine. The CAC was of particular interest as it determines the temperature leading into the combustion chamber, and this has a major impact of volumetric effciency of the cylinders and the rate of evaporation of the fuel, amongst others which influence the overall combustion performance.
pic1 pic2

The advantage of having a CAC which cools the air down to a lower temperature is that this air has a higher density, and so more molecules of air can be induced by the engine per cycle allowing for improved engine breathing. Graph 3 shows that, in order to achieve the same output power, for a poorer performing CAC, a higher pressure ratio over the compressor would be required. This higher pressure needed to pump this relatively hotter air would require more power from the turbine wheel, and so result in higher back pressure in the exhaust manifold – bad for scavenging, and less energy available for the MGU-H. The overall effective design of the cooling system has a holistic benefit on the F1 package, in so far as aerodynamics and power are concerned.
figure6
Low speed cooling issues
It is also important to consider the performance of the CAC at low vehicle velocities, as the cooling and pressure drop is highly dependant
of air velocity through the sidepods. A cooling map can be generated, which is a combination of the CFD and engine simulation work, which
shows that at lower vehicle speeds the car might not be cooling exactly as much as is needed. These maps can then be used to perform an energy audit of the total watts of heat rejected over the course of a lap and this will provide engineers feedback on the cars performance and that of the cooling system. Figure 6 shows a map for cooling Configuration 4 and Figure 7 is the heat rejection trace of around one lap of Circuit de Spa-Francorchamps.
table1
figure7

Relais und solid state Relais für Fahrzeugelektrik

[falloutboy]

Using Solid State Relays
Lets you do new things
by Julian Edgar

Relays have been around since the year dot, but solid state (ie electronic) relays are much newer devices. They don’t have any moving parts that can wear out, can be triggered by tiny currents but can in turn switch enormous currents, and can be tweaked to do quite a few interesting things.
Let’s take a look.

Traditional Relays
A relay is simply a remote-triggered switch. Traditional relays use a set of switching contacts and an electromagnet. When power is applied to the coil of the electromagnet, the contacts are physically pulled across, closing the switch.
fig01

Using Relays is a good grounding in relays and their automotive uses.
Traditional relays suffer from two problems. The first is that the more heavy duty a relay is (that is, the more current it can switch), the greater the current that’s drawn by the coil. This is because the switch contacts need to be larger (and so heavier), and the contacts need to be closed with force. These aspects require a strong electromagnet that draws higher currents.
The second disadvantage is that traditional relays wear out. The contacts tend to arc (especially when opening) and the contact faces get pitted. The more operations they perform, the faster they wear out.
However, a relay is a cheap and effective device, one which because of the variety of switch contacts that are available, can also be very versatile.

Electronic Relays
An electronic relay uses a big switching transistor to replace the mechanically moving contacts of a traditional relay. This means that an electronic relay will never wear out – nothing is moving. (Incidentally, the switching transistor is known as a MOSFET.)
fig02

In addition to never wearing out, an electronic relay can potentially switch very large currents. When equipped with a suitable heatsink, the relay shown here can handle 100 amps continuously and cope with a very short term switch-on current gulp of 240 amps. Those are huge numbers.
Electronic relays can be triggered with extremely small currents – for example, 20 milliamps. With traditional relays, only very light duty units have such low coil currents. The very low operating current has major advantages that we’ll come to in a moment.
Finally, an electronic relay can be pulsed very fast. Nope, we’re not talking about just switching on and off a few times a second, but instead being pulsed hundreds of times per second. In applications like motor speed control and light dimming, and injector/solenoid operation, this has major positives.

Jaycar SY4086
Das gleiche Relay das is dem Cosworth Artikel oben zu sehen war
he Jaycar SY4086 solid state relay costs about AUD$40. (Oh yes – that’s a disadvantage; electronic relays cost more than mechanical relays!) It’s made by Hongfa Relay and its data sheet can be downloaded by following the link at the end of this story.

The relay has four connections. The ones marked ‘input’ are connected to power (anything from 3 to 32 volts DC) to switch the relay on. Note that these connections are polarity conscious – positive must go to positive and negative to the car’s chassis or negative terminal of the battery.
The other end of the relay is the output – you can treat these terminals as the two connections of a switch (although note that the terminals are again polarised.)

Here is the relay wiring. In this diagram the relay is operated by a switch. Close the switch and power is applied to the left hand end of the relay – what in a mechanical relay would be the coil connections. This switches the relay on, so allowing current to flow to the load (the load could be a lamp, pump, motor, etc).
fig03

Here’s the same diagram but in this case the one power supply feeds the relay the juice needed to switch it on and also powers the load.
fig04

Finally, the electronic relay can be switched on and off by another electronic module – if the module outputs any voltage between 3-32V DC, it can be used to switch the electronic relay.
fig05

Uses

• Replacing Traditional Relays

The most obvious use of an electronic relay is to replace a traditional relay. However, in most automotive cases this isn’t worth doing. To replace existing relays, much rewiring will be need to be done and unless the existing relays are being used in an application that is causing them problems, no real benefits will occur.
However, there are some cases where it would still be a positive. For example, in a car using really high powered driving lights, and where the lights are switched on and off a lot, traditional relays may have a short life. The electronic relay would then be a good upgrade.

• Triggering From Small Currents

A better use is to use the electronic relay where only a small switch-on current is available.
For example, sensitive pressure switches, micro-switches and many temperature switches are rated for very low currents. But because of the low current draw of the electronic relay, these switches can be used without problems.
Existing switches such as radiator temperature switches and oil pressure switches can also be used without problems.
Pretty well any warning light or LED can be used to additionally trigger the relay – the relay’s input connections are simply wired in parallel with the LED or warning light (use a multimeter to check the polarity of the power feed to the light or LED.)

• Using an Electronic Module

If an electronic module has a pulsed output but cannot handle the load current, the electronic relay can be used to increase the capability of the module. For example, the The Nitrous Fuel Controller is a very cheap kit that can be used to:

• pulse injectors
• dim filament lights
• control the speed of motors
• pulse lights or horns

However, the output transistor of this module is limited to 10 amps. That’s a fair bit but not enough current to speed control big motors like radiator fans or fuel pumps, or happily pulse multiple car horns (eg as an alarm indication). But by using this module to trigger the electronic relay, each of these uses becomes possible.
However, note that there’s a trick to it. Rather than outputting 12V when the Nitrous Fuel Controller’s output is switched on, this module connects to ground when switched on. Therefore, the module connects to the ground line of the relay input rather than the positive side of the input.
(The new range of pre-built electronic modules that we’ll be covering soon will use the electronic relay in the same way – allowing the control of very large currents.)
fig06

• Adding a Delay

If more than a tiny current is available to trigger the relay, it’s very easy to add an extended ‘on’ time to the relay’s operation. This could be useful if for example an intercooler water spray is being operated by a pressbutton switch – you push the switch and release it and the spray continues for (say) 10 seconds.
fig07
To achieve the extended ‘on’ time, all that’s required is a capacitor wired across the input side of the relay. A 4700uF, 25V cap, for example, extends the ‘on’ time for about 7 seconds. To increase the delay, increase the capacitance. These capacitors are polarised, so make sure that the negative lead of the cap (shown by a line of minus symbols on its body) goes to the negative connection of the relay input and the positive side goes to the positive relay input. Cost of such a capacitor is only about $2.

Inductive Loads
There’s one other thing to keep in mind when using an electronic relay.
On inductive loads (eg those with coils like solenoids, injectors, motors and horns) a big voltage spike is produced when the power is turned off. To protect the relay, a diode (eg 1N4004) should be wired across the load, band on the diode towards the negative side of the relay output.

Heatsinking
When working at very high current levels, the electronic relay covered in this story requires heatsinking.
As this graph (taken from the data sheet) shows, at up to 40 degrees C ambient the relay can handle 25 amps continuously. At higher ambient temps (like found in an engine bay, for example!), the relay is de-rated.
If high currents need to be handled, bolt the relay to a large heatsink (using thermal grease between the heatsink and the metal back of the relay) and check that the relay does not exceed 40 degrees C in operation.
Fig08

Conclusion
In many cases traditional relays are still the best pick for the job. These applications include those where multiple switching contacts are required, cost is a factor and switched currents are small. But where large currents need to be handled, only very small currents are available to energise the relay, or where delayed ‘on’ times or high speed pulsing are required, the electronic relay is unbeatable.

Download the Jaycar SY4086 data sheet hereHFS33_en.pdf

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eigentlich sollte die PDF angehängt sein, aber da will wohl der Editor nicht so wie ich will

Vibration Analysis, Part 2: Driveline

[falloutboy]

Vibration Analysis, Part 2: Driveline
by GM

The sources of the vibration conditions that may be found on some 2014 Silverado 1500 and Sierra 1500 and 2015 Silverado, Tahoe, Suburban, Sierra, and Yukon models are most often the tires or driveline components, including axles and propeller shafts. These vibrations often occur at speeds of 35–45 mph (56–72 km/h) or 60–70 mph (96–120 km/h) and are felt in either the seat or steering wheel.

TIP: If the vibration can be duplicated on the rack, the test should be performed a second time with the wheel and tire assemblies removed from the vehicle and the wheel nuts installed to retain the brake discs and/or brake drums. If the vibration is eliminated, focus on the wheel and tire assemblies as the source of the vibration. If the vibration is still present, focus on the vehicle driveline as the source of the vibration.

Some vibrations may be difficult to diagnose even when the vibration can be duplicated. One example from a Technical Assistance Center (TAC) case was a concern about a vibration at 45 mph and higher on a 2014 Silverado 1500 4WD. The vibration was easily duplicated at these speeds. Initial diagnosis focused on a tire vibration.

A road test by a field service engineer using the CH-51450 Oscilloscope Diagnostic Kit with NVH showed a 1st order propshaft vibration with an amplitude of 7.83 mg at 49 mph. (Fig. 7)
fig07

The propshaft was balanced using the oscilloscope, but the condition did not improve.

TIP: For vehicles that are out of balance, perform a system balance. Using the two hose clamp method, the best driveline balance results are obtained under 10 g-cm. (Fig. 8)
fig08

Once the rear housing cover was removed, a 0.25–0.28 mm (0.010–0.011 in.) total variation of the backlash of the ring gear was found. The positions of the ring gear were swapped and side shims were installed to bring the backlash down to 0.1–0.12 mm (0.004–0.005 in.). However, the vibration was still present.

TIP: If the difference between all the measuring points is within specifications, the backlash at the minimum lash point measured should be 0.08–0.25?mm (0.003–0.010?in) with a preferred backlash of 0.13–0.18?mm (0.005–0.007?in).

The pinion and ring gear was replaced. A second road test showed a vibration amplitude of 0.722 mg at 45 mph. (Fig. 9)
fig09

Training and Tools
A recent TAC case on a 2013 Silverado 4WD illustrates the importance of proper training and the use of the correct tools during diagnosis.

In this case, after considerable time and multiple repairs, including a complete engine replacement, the vehicle was repurchased from the customer after the source of an engine idle vibration could not be found. The vehicle had a rough idle in gear during stops.

Using the CH-51450 Oscilloscope Diagnostic Kit with NVH showed a first order frequency with an amplitude of 5.66 mg at 525 RPM, which was significantly higher than a known good vehicle. (Fig. 10)
fig10

Once the baseline disturbance was measured, isolation of the first order engine disturbance and diagnosis could begin. Systematically, the serpentine belt was removed, and then the transmission torque converter was unbolted, which resulted in a first order engine frequency at an acceptable level. (Fig. 11)
fig11
The normal level of first order engine frequency was achieved by re-indexing the torque converter to the engine flywheel. No parts were required.

Vibration Analysis, Part 1: Tires

[falloutboy]

Vibration Analysis, Part 1: Tires
by GM

There have been several vibration conditions on 2014 Silverado 1500 and Sierra 1500 and 2015 Silverado, Tahoe, Suburban, Sierra, and Yukon models that have proved to be difficult to diagnose. These vibrations often occur at speeds of 35–45 mph (56–72 km/h) or 60–70 mph (96–120 km/h) and are felt in either the seat or steering wheel.

For example, one case from the Technical Assistance Center (TAC) was a concern about a vibration experienced at 70 mph and higher on a 2014 Silverado. At the dealership, after Road Force Balancing all tires, the Road Force Variation (RFV) measurements were: LF – 5 lbs., RF – 10 lbs., LR – 16 lbs., and RR – 20 lbs. In addition, the rear shocks were replaced.

A road test by a field service engineer using the CH-51450 Oscilloscope Diagnostic Kit with NVH showed a tire vibration amplitude of 22 mg at 70 mph. (Fig. 1) The tool’s sensor was placed on the seat track vertically.
fig01

Based on this information, the tires were moved from front to back on the same side. This put the highest RFV numbers on the front and the lowest on the rear of the vehicle. A second road test showed a greatly reduced tire vibration amplitude of 0.804 mg at 69 mph. (Fig. 2)
fig02

Some of the vibration cases may be difficult – but with the right approach and the right tools, a successful diagnosis can be achieved quickly.

Diagnostic Information
A variety of helpful information is available in the Service Information.
Information on Vibration Analysis and Diagnostics – #PI1354B provides detailed information on vibration analysis and diagnosis for several different conditions. It outlines the recommendations and procedures for diagnosing and repairing vibrations caused by wheels and tires, axle components and propeller shafts. It also includes a vibration diagnostic worksheet to record vibration measurements.

Vibration Analysis Worksheet – Bulletin 03-00-91-001G is a vibration analysis worksheet that is to be used when road testing vehicles exhibiting vibration conditions. The worksheet lists the necessary data needed in conjunction with the appropropiate testing procedures in the Service Information.

Information Needed when Calling TAC – When calling TAC for diagnostic help on vibration conditions, there is some basic diagnostic information needed in order to provide proper direction in repairing a vehicle. Before calling, technicians should use #PIT5345 to understand what is needed on a vibration condition. The measurements listed in the PI should be gathered using the appropriate tools.

First Steps

The first step in determining the cause of a vibration is a test drive with the appropriate diagnostic equipment installed on the vehicle. If the correct tools are not used or the proper procedures are not followed, an incorrect diagnosis will result.

• Inspect the truck for any aftermarket equipment installations, such as non-factory tires, lift kits or running boards.
• Mark each tire valve stem location on the tire to check for tire slippage on the rim. After the road test, verify that the tires have not slipped on the rim.
• Use the CH-51450 Oscilloscope Diagnostic Kit with NVH for vibration diagnosis. The oscilloscope kit provides an accurate analysis of vehicle noise, vibration and harshness conditions and uses the display of your laptop computer to present clear, easy-to-read results and actions for repair.
• For seat vibrations, mount the oscilloscope kit sensor to the rear seat bracket. (Fig. 3, #1)
• For steering wheel vibrations, mount the oscilloscope kit sensor to the steering wheel bracket under the instrument panel. (Fig 3, #2)
• Moving the oscilloscope kit sensor from a vertical position to a horizontal position may indicate higher amplitude, which may be beneficial to help in diagnosis.
• Measure the vibration by driving the truck with the transmission in M5 to keep the vehicle from switching in and out of Active Fuel Management (AFM).
• Once the vibration readings have been recorded on a road test, verify the vibration data in the service bay. If using a hoist, the suspension must be at the same trim height as when the vehicle normally sits on the road.

fig03

TIP: If the vibration can be duplicated on the rack, the test should be performed a second time with the wheel and tire assemblies removed from the vehicle and the wheel nuts installed to retain the brake discs and/or brake drums. If the vibration is eliminated, focus on the wheel and tire assemblies as the source of the vibration. If the vibration is still present, focus on the vehicle driveline as the source of the vibration.

Another case example from TAC shows the importance of understanding how to use the right tools. The case was a vibration condition on a 2015 Tahoe felt at 40-50 mph and at 60-70 mph.

Four different tires, a rear driveshaft and a rear differential assembly had been installed to address the vibration condition.

After a road test, the field service engineer determined the primary vibration was a 2nd order tire disturbance.

Tire vibrations were measured with the CH-51450 Oscilloscope Diagnostic Kit. The initial RFV measurements for three tires were 25 lbs., 16 lbs., and 12 lbs. The fourth tire had a measurement of 8 lbs. (1st order disturbance), but also a 2nd order disturbance of 21 lbs. (Fig. 4)
fig04

Reviewing the Hunter GSP9700 Road Force Balancer results for the tire with the 2nd order disturbance showed the 1st order harmonic was below specification, but the 2nd order specification was 21 lbs. (Fig. 5) It’s important to look at all harmonic measurements when reviewing the road force measurements and not to dismiss a particular tire based on only one measurement. If present, the CH-51450 Oscilloscope Diagnostic Kit tool will display the primary vibration as a 2nd order disturbance. Be sure to use this information and look at the other harmonic measurements on the Hunter GSP9700 Road Force Balancer.
fig05

The vibration was corrected by replacing and match-mounting (or vectoring) all four tires. The RFV measurements were 1 lb., 4 lbs., 4 lbs., and 7 lbs. (Fig. 6)

#PI1354B lists a RFV specification of 15 lbs. for light truck tires. This specification is lower than the current specification listed in the Service Information. It should only be used if there is a speed-related tire disturbance. If the RFV of the tire is over the specification, match-mount (vector) the tire on the wheel. If that doesn’t bring down the measurement to within the specification, the tire should be replaced.
fig06

TIP: When replacing tires, the road force should be checked before a test drive and after a test drive (minimum of 10–15 miles or 16–24 km/h). Road force on new tires will change dramatically after the tires are warmed up (as much as a 20-lb. reduction). After the test drive, the tire’s road force should be checked. If acceptable RFV cannot be achieved, first try vectoring the tire on the rim before an alternate tire is used. Refer to Bulletin #13-03-10-002: Diagnostic Tips for Difficult to Resolve Tire/Wheel Vibration Concerns.