ROTEC Publications - 81 - rotec In Figure 3 the valve lift, velocity and acceleration are shown for various camshaft speeds. Before and after maximum valve lift, effects due to valve spring oscillations may be seen in both the velocity and acceleration curves. Acceleration data can be directly related to valve loads which peak before and after valve seating. Over the range of valve seating, an unwanted phenomenon appears – valve bounce. This valve bouncing phenomenon will accelerate seat wear, results in a loss of power and may cause engine damage. In the valve closing range, valve bounce can be observed in the valve lift, valve velocity and acceleration profiles. The peaks which appear in these curves are due to this phenomenon. Figure 3: Valve lift, valve velocity and valve acceleration at different speed values Valve Lift 60 90 120 150 180 210 240 270 300 Angle Camshaft [deg] -2 0 2 4 6 8 10 Lift [mm] 1000 rpm 2000 rpm 2500 rpm 2700 rpm 2900 rpm Valve Velocity 60 90 120 150 180 210 240 270 300 Angle Camshaft [deg] -0.020 -0.015 -0.010 -0.005 0 0.005 0.010 0.015 0.020 norm. Velocity [m/rad] 1000 rpm 2000 rpm 2500 rpm 2700 rpm 2900 rpm Valve Acceleration 60 90 120 150 180 210 240 270 300 Angle Camshaft [deg] -0.2 -0.1 0 0.1 0.2 0.3 0.4 norm. Acceleration [m/rad²] 1000 rpm 2000 rpm 2500 rpm 2700 rpm 2900 rpm
Transcript
ROTEC Publications - 81 -
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In Figure 3 the valve lift, velocity and acceleration are shown for various camshaft speeds. Before and after maximum valve lift, effects due to valve spring oscillations may be seen in both the velocity and acceleration curves. Acceleration data can be directly related to valve loads which peak before and after valve seating. Over the range of valve seating, an unwanted phenomenon appears – valve bounce. This valve bouncing phenomenon will accelerate seat wear, results in a loss of power and may cause engine damage. In the valve closing range, valve bounce can be observed in the valve lift, valve velocity and acceleration profiles. The peaks which appear in these curves are due to this phenomenon.
Figure 3: Valve lift, valve velocity and valve acceleration at different speed values
Figures 4 and 5 show two further standard analyses in the ROTEC software:
- 2D plots of valve lift, velocity and acceleration at selected speed values (Figure 4)- Plots of valve closing velocity and valve closing angle at a specified value of lift in the valve closing range (Figure 5)
Figure 4: Valve lift, velocity and acceleration at 1000 and 2900rpm
Figure 5: Valve closing velocity and closing angle at 0.1mm lift versus speed
Valve Closing Velocity and Closing Angle at 0.1mm Lift
500 1000 1500 2000 2500 3000
Speed Camshaft [rpm]
-1.6
-1.4
-1.2
-1.0
-0.8
-0.6
-0.4
-0.2
0
(1) V
eloc
ity [m
/s]
230
231
232
233
234
235
236
237
238
(2) Angle Cam
shaft [deg]
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Measured lift curves versus speed may be compared with either 2D theoretical lift curves or reference lift curves measured at low speeds. These comparisons highlight another dynamic valve phenomenon – valve jump at high engine speed (Figure 6). Under normal conditions, the valvetrain components including the valve, tappet, spring retainer etc. are pressed against the cam surface by the valve spring and move up and down while maintaining contact with the cam lobe as the camshaft rotates. However, due to high acceleration the valve can unexpected lift up and lose contact with the cam profile. Besides poor NVH behaviour this can cause wear (pitting) on the cam profile where the valve clashes back onto the cam (closing ramp). In severe cases, heavy lift overshoot can result in valve-piston contact leading to engine damage.
Figure 6: 3D difference lift plot versus speed (this results from subtraction of a 2D single lift cycle measured at low speed
Figure 7 shows a speed ramp performed on another cylinder head test bench powered by an electric motor. This test bench is designed to simulate a 4-cylinder engine. In this case the camshaft torque and valve spring forces were also measured. The data may be used in investigating valve spring dynamics. Frequency analysis of both the spring force and torque curves can reveal spring natural frequencies typically in the range of 400 to 500Hz.
Figure 7: Speed ramp from 500 to 2500rpm camshaft speed
Figure 8: 3D waterfall plot of valve spring force
Figure 9: 3D waterfall plot of camshaft torque
Camshaft Speed [rpm]
0 20 40 60 80 100 120 140 160
Time [s]0
500
1000
1500
2000
2500
3000
Spee
d [rp
m]
Valve Spring Force
0 60 120 180 240 300Camshaft Angle [deg] 0
500
1000
1500
2000
2500
3000
Speed [rpm]-200
0
200
400
600
800
1000
1200
Forc
e [N
]
Camshaft Torque
0 60 120 180 240 300Camshaft Angle [deg] 0
500
1000
1500
2000
2500
3000
Speed [rpm]-10
0
10
20
Torq
ue [N
m]
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Figure 10: Valve lift, camshaft torque and spring force at 2500rpm camshaft speed
2.3.2 VALVETRAIN TESTING ON FIRED ENGINES
Compared to motored test benches, valvetrain testing on running engines reflects real working conditions and, in particular, temperature effects. Valve displacement can be measured using both specially designed transducers and commercially available sensors. Apart from the difficulties encountered in mounting the sensors they are often limited in the range of valve motion which can be measured. In recent years magnetoresistive sensors mounted inside the cylinder in the valve guide have been successfully used for measuring valve lift [7, 9, 10]. Major advantages of these sensors include small installation space requirements and operation at high temperatures. In order to measure valve displacement the sensor requires a (ferromagnetic) tooth structure machined into the valve stem. Movement of the valve, and thus the teeth, produces a change in the sensor’s electrical resistance due to the varying magnetic field. The sensor outputs simultaneous sine and cosine signals resulting from the valve movement. These signals are input to ROTEC 400kHz analogue channels. A ROTEC software algorithm generates a valve lift curve from the sine and cosine signals.
The accuracy of this valve lift curve may be determined by performing measurements with both the magnetoresistive sensor and high speed vibrometer in parallel on a motored test bench before the machined valve, sensor assembly, etc. are used on the fired engine. Figure 11 shows the measured magnetoresistive sensor and laser signals while Figure 12 shows the result of the comparison where the valve lift calculated from the magneto-resistive sensor signals has been subtracted from the lift curve measured with the laser. The maximum deviation is of the order of 20 microns which is acceptable for all practical purposes.
Valve Lift, Camshaft Torque and Spring Force at 2500rpm
0 60 120 180 240 300 360
Camshaft Angle [deg]
-200
0
200
400
600
800
1000
1200
(1) F
orce
[N]
-15
-10
-5
0
5
10
15
20
(2) Torque [Nm
]
-2
0
2
4
6
8
10
(3) Lift [mm
]
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Figure 11: MR sensor sine and cosine voltage outputs and valve lift signal (HSV laser)
Figure 12: Measured lift difference between HSV laser and magnetoresistive sensor
Magnetoresistive Sensor and Laser Signals
60 120 180 240 300
Camshaft Angle [deg]0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
4.0
(1) V
olta
ge [V
]
0
2
4
6
8
10
12
(2) Valve Lift [mm
]
Valve Lift Calibration Curve [mm]
7 9Camshaft Revolutions-2
0
2
4
6
8
10
12
(1)
Valv
e Li
ft [m
m]
-0.010
-0.005
0
0.005
0.010
0.015
0.020
0.025
(2) Difference [m
m]
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1. DUAL MASS FLYWHEEL TESTING
Another frequent application for ROTEC equipment is in-vehicle testing of clutches and dual mass flywheels (DMF). To identify dynamic torsional vibration effects in torque transmission, simultaneous measurement of speed on both sides of the clutch or DMF is required. On the engine side, the starter ring gear may be scanned with a magnetic sensor. On the transmission side, a gear within the gearbox must be accessed or, alternatively, a toothed rim can be attached to the gearbox input shaft. This angular sampling provides a fixed number of data points per revolution which is independent of the rotational speed. The momentary angular velocity of rotating shafts is thus measured, i.e. the mean velocity from pulse to pulse. The vibration angle and the angular acceleration are obtained by integrating and differentiating the measured angular velocity respectively. These two calculations are important when investigating torsional vibration problems. Another important calculation is the angle between the two speed channels (angular displacement, slip).
3.1 CLUTCH MEASUREMENTS
Figure 13 shows test results from a speed run-up in fourth gear (4-cylinder engine, conventional clutch). Magnetic speed sensors were used on the starter gear (132 teeth) and on a gearwheel on the gearbox input shaft (27 teeth).
Figure 13: Speed ramp. Time history data on both sides of the clutch
Figure 14 shows two revolutions of the speed curves in detail. The data points from the speed measurements are also shown. The engine‘s firing order can be clearly seen (2nd order, twice per revolution).
An FFT analysis shows additional orders since cyclical combustion is not a purely sinusoidal process. The 3D waterfall plots (Figure 15) show order analyses of angular acceleration as a function of speed. The gearbox acceleration peaks which occur at resonance-critical speeds contribute, for example, to the unwelcome audible noise known as gearbox rattle.
0 5 10 15 20
Time [s]
500
1000
1500
2000
2500
3000
3500
4000
4500
Spee
d [rp
m] Engine
Gearbox
3
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These 2nd order resonance peaks can be reduced by modifying clutch friction and torsional stiffness. Alternatively, a dual-mass flywheel (DMF) may be used. The main purpose of a DMF is to counter gearbox rattle by damping the input torsional vibrations.
Figure 14: Speed data over two revolutions
Figure 15: Angular acceleration [rad/s2], engine and gearbox
200.0 200.5 201.0 201.5 202.0Engine Revolutions
2150
2200
2250
2300
2350
2400
2450
Spee
d [rp
m]
Gearbox
Engine side
Engine
0
1
2
3
4Order
10001500
20002500
30003500
4000
Engine Speed [rpm]
0
500
1000
1500
2000
2500
3000
3500
4000
4500
5000
Angu
lar A
ccel
erat
ion
[rad/
(s*s
)]
1: 2.0000
Gearbox
0
1
2
3
4Order
10001500
20002500
30003500
4000
Engine Speed [rpm]
0
1000
2000
3000
4000
5000
Angu
lar A
ccel
erat
ion
[rad/
(s*s
)]
2: 2.0000
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Figure 16: 2nd order angular acceleration [rad/s2], engine and gearbox
3.2 DMF MEASUREMENTS
Dual mass flywheels (DMF) are installed in many modern vehicles with manual transmissions. The DMF is located between the engine and the gearbox replacing the conventional flywheel. The DMF isolates the driveline from engine excitation thus increasing driving comfort. The primary inertial mass is connected to the output shaft of the engine. The secondary inertial mass is connected to the input shaft of the gearbox thus increasing the moment of inertia of the transmission. These two de-coupled masses are linked by an elastic spring system which permits relative rotary motion between the primary and secondary masses (Figure 17). The use of the secondary mass and appropriate spring constants have the effect of shifting the resonance speeds which excite the natural rattling frequency below the engine idling speed, i.e. outside the normal driving ranges. The secondary mass also forms one of the two friction surfaces for the clutch disc.
Figure 17: Dual-mass-flywheel, schematic assembly
Angular Acceleration, Engine and Gearbox
500 1000 1500 2000 2500 3000 3500 4000 4500
Engine Speed [rpm]
0
1000
2000
3000
4000
5000
Angu
lar A
ccel
erat
ion,
pea
k [ra
d/(s
*s)]
Primary mass
Secondary mass
Starter gear
Spring damping system
ClutchGearbox
Engine
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The main disadvantage of the DMF is that every time the engine is started or stopped it has to run through this resonance point. Problems can then arise if e.g. the vehicle gets stuck in resonance. A further disadvantage of the DMF is an increase in engine-side torsional vibrations due to the DMF‘s lower effective moment of inertia on the engine side compared to a conventional flywheel. Figure 18 shows time history rotational speed data taken on a diesel engine both with and without a dual-mass-flywheel. The measurements were performed at idle speed (flywheel and gearbox input). The DMF causes an increase in engine-side vibrations (red curve).
Figure 18: Speed data with and without a DMF
The main disadvantage of shifting transmission resonance below idle speed is that when the vehicle is started it has to run through resonace. Figure 19 shows data from a normal engine start – the engine runs through the resonance point and reaches idle speed. The DMF angular displacement reaches a maximum of about 45 degrees.
Figure 19: Good engine start behaviour. Time and angular displacement
0 0.05 0.10 0.15 0.20 0.25 0.30
Time [s]
-125
-100
-75
-50
-25
0
25
50
75
Spee
d Fl
uctu
atio
n [rp
m]
Flywheel, no DMF Flywheel, with DMS Gearbox, no DMF Gearbox, with DMF
1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0Time [s]
0
100
200
300
400
500
600
700
800
900
1000
1100
1200
(1) S
peed
[rpm
]
-120
-100
-80
-60
-40
-20
0
20
40
60
80
100
120
(2) Relative Angular D
isplacement [degrees]
Engine side
Gearbox side
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Figure 20 shows data from a bad engine start. The magnitude of the DMF angular displacement causes the spring components to lock-out. The maximum DMF angular displacement is 60 degrees in both directions. The resulting torque causes the friction clutch disc to slip through and the angular displacement curve runs away. In this case, poor engine management parameters (injection volume, duration and time) in combination with a high DMF spring rate caused the bad engine starting behaviour. The aim of this type of measurement is to optimise the starting conditions and keep the DMF angular displacement below a set value (for this particular DMF below 45 degrees).
Figure 20: Bad engine-start behaviour. Time and angular displacement curves
Figures 21 and 22 show a case of deliberate vehicle abuse (total misuse duration of around one minute). The vehicle is sharply braked in 4th gear until drive train vibrations arise below idle speed. Figure 21 shows the speed data. ROTEC speed sensors with directional recognition were used (starter gear with 132 teeth, toothed rim with 170 teeth attached to the secondary side). The data show large speed fluctuations when resonance is reached and the direction of rotation reverses for a short time.
Figure 21: DMF driveline resonance. Time histories
4.0 4.5 5.0 5.5 6.0 6.5 7.0 7.5Time [s]
0
100
200
300
400
500
600
700
800
900
1000
1100
1200
(1) S
peed
[rpm
]
-120
-100
-80
-60
-40
-20
0
20
40
60
80
100
120
(2) Relative Angular D
isplacement [degrees]
Gearbox side
Engine side
5 6 7 8 9 10 11 12 13 14 15 16
Time [s]
-600
-400
-200
0
200
400
600
800
1000
1200
Spee
d [rp
m]
Gearbox side
Engine side
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With decreasing engine speed, the excitation in the transmission increases until, at resonance, the primary to secondary mass angle reaches 120 degrees and the maximum permissible engine torque is exceeded (Figure 22). This can damage the DMF or other driveline components. Torque overload protection is provided by a torque limiter in the DMF. When the torque limit is exceeded the DMF slips between the primary and secondary masses. The drivetrain remains in resonance unless the clutch is activated or the vehicle is brought to rest.
The author wishes to thank Dirk Beismann of Ford-Werke GmbH in Köln for useful discussions on the valvetrain data presented in Section 2 and Ralf Till of ZF AG, Schweinfurt for providing the measurement data presented in Section 3.
