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SPRINGER BRIEFS IN
APPLIED SCIENCES AND TECHNOLOGY
Taehun Seung
Self-Induced Fault
of a Hydraulic
Servo Valve
A Possible Cause for
Hidden Malfunction
of Aircraft’s Systems
SpringerBriefs in Applied Sciences
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Indexed by EI-Compendex, SCOPUS and Springerlink.
Self-Induced Fault
of a Hydraulic Servo Valve
A Possible Cause for Hidden Malfunction
of Aircraft’s Systems
123
Taehun Seung
Ockenfels, Germany
This Springer imprint is published by the registered company Springer Nature Switzerland AG
The registered company address is: Gewerbestrasse 11, 6330 Cham, Switzerland
Acknowledgements
The author would like to thank his colleagues at his former employer Mr. Thomas
Fricker for the valuable contribution whilst troubleshooting and Mr. Klaus Baldauf
and Mr. Rudolf Scheiblich for their indefatigable supports whilst laboratory
investigation.
The same is valid for his colleagues at the partner companies in USA and Brazil.
The professional advice and proof-reading of Ms. Jane Lawson in writing this
publication are also gratefully acknowledged.
v
Contents
vii
viii Contents
ix
Introduction
xi
xii Introduction
Due to this, the closed loop control is no longer able to manage the system. This
interaction is more complex than ‘sticking’ of a spool and sleeve assembly.
This report1 details the investigation and result made recently in the field of
aircraft engineering and introduces the entire working mechanism of a self-induced
fault of a servo valve.
1
This is the final comprehensive description a shortened congress paper for 31st ICAS originates
from. The paper ‘ICAS 2018_0853’ was presented in Belo Horizonte, Brazil, on 10 September 2018.
cf. https://www.icas.org/ICAS_ARCHIVE/ICAS2018/data/papers/ICAS2018_0853_paper.pdf
Topics and Arrangement of This Report
xiii
Chapter 1
Occurrence and Suspicion
Brake Control
Shut-off valve
Valve
P P
1 4 2 3
R R
P P P P
Hydraulic Hydraulic
System 1 System 2
R B R B R B R B
Pressure
Transducer
PT PT PT PT
Wheel Speed
Transducer
W 1 2 W W 3 4 W
Pressure
Accumulator
Switch
Wheel
Check valve Shuttle Valve
P BB P
R R
Hydraulic Hydraulic
System 1 Dual Emergency / System 2
Park Brake Valve
R-Nozzle P-Nozzle
Flapper
R P
C
Spool & Sleeve (Second stage)
R P
Spool
B Brake pressure
in a separate housing. The principle schematic is shown in Fig. 1.2. The primary
stage employs no cantilever feedback spring, i.e. there is no mechanical corre-
sponding connection between primary and secondary stages (cf. Sect. 3.2.2 Ref.
[13]). A coil spring in the secondary stage resets the spool whenever the flapper
reduces the control pressure to the secondary stage. It must be said that the reset
mechanism has hardly any effect on the circumstances and phenomena described
hereafter. Furthermore, the physical phenomena and resulting fault mechanism
described below can occur in any kind of pilot-working hydraulic servo valves.
Figure 1.3 shows the relation of ‘pressure response versus spool stroke’ in the
secondary stage. The plots are made at a low, quasi-static demanding condition for
which the spool was moving less than 0.16 mm/s.
Having passed a certain overlapped range at the beginning (i.e. 0–0.485 mm),
the pressure answer corresponds directly proportional to the spool’s stroke and
consequently to the geometrical opening gap given between the edges of the spool
and sleeve (i.e. 0.493–0.753 mm).
In contrast to the plots of quasi-static demanding shown above, Figs. 1.4, 1.5,
1.6 and 1.7 show some characteristic curves of the servo valve under different
dynamic demanding conditions for which the demanding speed has been changed.
It is easy to recognize that the hysteresis of the characteristic curve becomes more
violent by increasing the demanding speed.
Regardless the demanding speed, i.e. despite quasi-static or dynamic demanding
conditions, the spool adjusts at the actual working point to establish an equilibrium
state, whereas the magnetic field strength and consequently the opening rate of the
4 1 Occurrence and Suspicion
Fig. 1.3 Spool stroke versus Pressure response in the second stage
Fig. 1.4 Characteristic curve of the servo valve at a demanding speed of 1 mA/s
1.2 Appraisal and Fundamental Contemplation … 5
Fig. 1.5 Characteristic curve of the servo valve at a demanding speed of 4 mA/s
Fig. 1.6 Characteristic curve of the servo valve at a demanding speed of 24 mA/s
6 1 Occurrence and Suspicion
Fig. 1.7 Characteristic curve of the servo valve at a demanding speed of 100 mA/s
inlet nozzle in the first stage, friction/lubrication between spool and sleeve, internal
leakage flow, the strength of the return spring, the inertia of the spool, etc., are the
main parameters (cf. Sect. 3.1.1.1).
After having carried out extensive test campaigns with stepwise increased
contamination rate, some experts have even concluded that the servo valves are
very much more insensible than they are usually supposed to be [1]. A similar
laboratory test campaign carried out for this actual investigation confirmed this.
Moreover, despite sufficient filtration of the fluid, the occurrences mentioned in
Sect. 1.1 were still recurring with nearly the same fault rate. So far, debris was
excluded from the list of possible root causes for the present investigation.
As stated in the introduction, the present work deals solely with uncontrollability
caused by purely fluid mechanical occurrences, i.e. non-debris-faults.
In most cases, the torque motor consists of an electric coil and relatively simple
(preloaded) mechanical moving part which stands under the effect of one or more
permanent magnets. The electric coil can be a simplex or duplex type and creates/
regulates the necessary magnetic field by demanding in order to control the initial
movement of the (pilot-working) valve system.
Dropout of such electromagnetic parts due to any kind of interference, like
foreign magnetic field, vibration/shaking and/or shock, is conceivable in spite of
shielding/insulation. But this possibility is not to be taken into consideration already
as a root cause since there was no typical error pattern. In the case of the present
investigation, the fault still occurred at many serial units without changing in fault
rate, despite extremely varying electromagnetic working conditions. The fault still
occurred even in an insulated environmental condition (both vibration and magnetic
field). Hence, it was uncertain that the fault had been caused by foreign influences
—such as debris.
In terms of movement ability of a spool, its (radial) fit tolerance relative to the sleeve
is of prime importance. It is even trivial to mention that the effective fit tolerance will
be determined by the actual temperature. The development of modern numeric
simulation tools has been made quite good progress during the last decades, and they
are able to predict the possible thermal effects with a high level of accuracy. Thermal
effects, therefore, can be considered as an underpart with minor importance, as long
as this does not give rise to mutual effects with other parameters.
8 1 Occurrence and Suspicion
Besides the radial fit tolerance of a spool and sleeve assembly, the actual angular
position of the spool relative to the sleeve could have a significant influence on the
movement ability of the spool. This comes from the straightness tolerances of both
spool and sleeve. In fact, neither the spool nor the sleeve is absolutely straight and
free of bends since it is impossible to machine such a perfect geometrical shape.
Should both members be in paraphrase with their bends at an indeterminate
moment, i.e. convex to concave, the actual friction of the sliding surface would be
maximized and vice versa (cf. Fig. 3.12). Note that this is only a reflection of a
two-dimensional case as a simplified thought experiment. In reality, the curve
progressions are three dimensional. In any case, the friction between the spool and
sleeve changes in accordance with the actual angular position of the spool. Hence,
the straightness of the members can be of decisive parameter and particular
importance, even though this often is taken as a matter of course.
In order to investigate the possible influence of the spool’s angular position on
the pressure development in the characteristic diagram, a test campaign was made.
The spool was prepared with a mark so that different angular positions could be
set relative to the sleeve (cf. Fig. 1.8. Note that the spool is actually set to the 12
o’clock position.). The test was performed with four different spool positions preset
between the test campaigns.
Whilst field operating, the spool is able to roll inside the sleeve an indeterminate
amount of angle and direction corresponding to the actual dynamic pressure, flow
rate, leakage stream, running speed of the spool, etc. In order to keep the preset
position of the spool, no consumer was connected to the valve. The flow rate, which
is one of the main parameter, was eliminated during the whole test in this way. The
test was conducted at a standardized procedure at which the running time and pause
were kept same.
The measurement results plotted in Figs. 1.9, 1.10, 1.11 and 1.12 show clearly
the significant influence of the spool’s angular position on the pressure development
(Every plot contains ten cycles demanded from 2.647 to 57.533 mA at a constant
demanding speed of 24 mA/s.).
The curve runs shown in Figs. 1.9, 1.10 and 1.11 strew in a certain range and
even change its gradient at each cycle. In contrast, the curve runs shown in
Fig. 1.12 are strikingly smooth and keep its gradient throughout the whole series of
cycles.
Comparing the diagrams with that in Fig. 1.12, it is concluded that the running
speed of the spool changes and it is dependent on the actual angular position of the
spool. Moreover, it is recognized that the sliding ability at the same point is getting
worse with increasing number of cycles (cf. Figs. 1.9, 1.10 and 1.11). The reason
for such changing in sliding ability seems to be the degeneration of the lubrication
film caused by rubbing (cf. Sect. 3.1.3.1).
Note that Schlemmer et al. [12] carried out a similar investigation and reported a
certain angle dependency of the running friction of the spool.
The balancing grooves on the spool are occasionally called ‘relief grooves’, and
there are numerous detailed studies and suggestions in terms of form, number and
10 1 Occurrence and Suspicion
distance between each balancing groove [2, 3, 7–12, 14, 15]. In order to make the
function and working mechanism of the balancing grooves clear, simplified qual-
itative contemplations will be briefly discussed: A balancing groove is able to work
correctly only when the groove is able to build a closed annular chamber around the
spool.
In other words, the functionality of balancing grooves may only be guaranteed
when they are covered by the sleeve. Figure 1.13 shows the principle and func-
tionality of such an annular pressure chamber; the fluid from the higher pressure
side tries to compensate its pressure difference by leaking to the lower pressure side.
During the process, the leakage stream will try to find a way of lowest resistance.
Hence, the leakage stream takes the route with the largest geometric gap. This does
not necessarily have to be a direct route. In the reality, the route can be slightly bent
in a radial direction so that it looks like a spiral line. Without having a working
balancing groove, the pressure will be decreased proportionally among the
streamline (cf. Fig. 1.14). Doing that, the pressure force would press the spool to
the sleeve wall. The gap between the spool and sleeve could be increased as the
spool’s centreline is no longer aligned coaxially with that of the sleeve. Once a
small misalignment occurs, the situation will become worse and worse at every
movement. The spool is not able to recover the misalignment as the side force will
be radically increased.
12 1 Occurrence and Suspicion
Sleeve
Pressure
Vektor sum
Pressure profile
Length
Fig. 1.13 Abolition of infinitesimal radial forces at perfect working balancing grooves as the
result of the coaxial alignment of the spool relative to the sleeve
Sleeve
Pressure
Vektor sum
Pressure profile
Length
Should a balancing groove exist and still be working, the local pressure differ-
ence around the spool will be balanced since the local pressure around the spool’s
circumference changes thanks to the closed annual chamber equally and simulta-
neously [5]. During this process, the spool recovers a possible coaxial misalignment
by itself (cf. Fig. 1.13). In other words, this means that the spool ‘relieves’ by itself.
As a result, the leakage stream in circumference of the spool is evenly dispensed
and consequently the flow rate, i.e. leakage, is timely constant (cf. Sect. 2.1.2).
Excluding debris and dropout of the electric parts as possible root causes (cf.
Sects. 1.2.3.1 and 1.2.3.2), there are hardly any other fault-provoking parameters
and/or problem potentials than friction on the sliding surface area of the spool and
sleeve assembly.
This kind of mechanical problem known as ‘hydraulic locking’ has existed for
ages, practically since the spool and sleeve assemblies came into use for hydraulic
control devices.
As briefly mentioned in Sect. 1.1, this traditional terminology is a bit of a
misnomer, in so far as it is used for the spool and sleeve assembly. It expresses the
situation as if the valve were blocked fast and durably. In reality, the spool sticks
only temporarily, so far no deformation has occurred at a member (i.e. either spool
or sleeve) or a certain situation freezes the situation (cf. Sect. 1.3). Some parametric
studies about such ‘sticking’ of a spool and sleeve assembly were made, and
countermeasures have been developed [2–5, 7–12, 14, 15].
It is fully impracticable and undesirable to assemble the spool and sleeve com-
pletely without internal leakage unless the motion of the spool is managed by high
actuation force, e.g. direct drive. Having an adequate amount of internal leakage,
the spool is able to ‘swim’ inside sleeve. The leakage stream in the gap between the
spool and sleeve flushes possible impurity, and above all the fluid functions as a
lubricant. The prerequisite, however, is an even gap distance throughout the whole
14 1 Occurrence and Suspicion
h1 h2
h2 ≈ 2∙h1
Sleeve S1 = S2
Fig. 1.15 Change of the gap distance despite same opening surface area
where
h: gap height, p: pressure, U: velocity in x-direction
x: tangential coordination, W: velocity in z-direction
z: tangential coordination, η: dynamic viscosity
Figure 1.15 shows the geometric condition in the case of a metallic contact due
to a non-coaxial misalignment of the sleeve. On the opposite side of the contact
point, the gap amount increases to double compared to the original gap distance and
consequently the internal leakage, in other words the ‘purging quantity’, increases
abruptly according to the equation given above. Once a continuous axial flow is
established, the spool is hardly able to recover the original gap by itself because the
resulting lateral force of the spool presses the spool to the sleeve wall (cf.
Fig. 1.14).
In the case of a radial contact of the spool to the sleeve’s wall, it is also to expect
that the friction will be drastically increased.
Air in the hydraulic circuit is not only as a result of an insufficient bleeding process
but also accrues in fact as a ‘generative’ problem. The amount of air in the hydraulic
1.2 Appraisal and Fundamental Contemplation … 15
fluid increases during the operation as the mechanical parts, which are shuttling two
medium zones, like the piston rod, will continue to bring new gas molecules into
the fluid side at every stroke.
Vacuum accelerates the ‘sliding-surface degeneration’ process: due to a variety
of reasons, the pressure of a hydraulic circuit could sink down locally even below
atmospheric pressure during the operation. In such a case, the gas molecules
absorbed in the fluid can suddenly escape and create bubbles.
A detailed analysis was made during the troubleshooting phase of the current
investigation. Once bubbles are created, significant ‘sponge effect’ was the con-
sequent result at least in the following command cycle in the case of a
brake-by-wire control system.
Generally, the first stage is less sensitive than the second stage even if it has more
parts and requires a fine adjustment in most cases. Once assembled properly and
protected by an adequate ‘last chance filter’, shielded against electromagnetic
effects and protected against possible vibrations, the first stage is rather robust. This,
however, does not necessarily mean that the flow control mechanism of the first
stage is immune to any kind of instability. The system, for example, can react
sensitively to fluid mechanical phenomena: should the control system accidentally
work across a working point, at which the flow separation occurs, the effect of such
a transition must be noticeable in the electro-hydraulic signal conversion (unique
incidence working point).
By reaching the sub-critical flow region (103 < Re < 1.7 * 4 105), the flow
separates on the cylindrical flapper body, so that this experiences an extra drag. The
flapper drifts away to the downstream direction as if a sudden suction force would
have been developed on its lee side. Whenever such flow separation occurs on the
flapper body, the gap at the nozzle increases accordingly a certain amount due to the
extra force.
The effect of such an extra opening at the nozzle clearly reflects in the pressure
answer as shown in Fig. 1.16 as a sudden increase in the gradient (see Sect. 2.1.1
for the pressure sensor positions). The extra opening of the nozzle will be kept until
the reversal of the command—as long as the system is able to create a sufficient
reset force afterwards to manage the flapper movement (cf. Sect. 3.1.2).
It must be said that such a working point is difficult to detect in the reality due to
numerous, simultaneously changing nonlinear parameters, even though there is no
doubt of its existence.
Figure 1.17 shows the corresponding time history of Fig. 1.16. It is easy to
recognize that the pressure answer sporadically inclines to override at a preset
current limitation of 19.8 (mA).
16 1 Occurrence and Suspicion
Fig. 1.16 Sporadic sudden increase of the pressure due to extra suction force at the flapper
Fig. 1.17 Sporadic override of the pressure answer despite the current limitation
1.2 Appraisal and Fundamental Contemplation … 17
Turbulence Flapper
P-Nozzle
(Inlet-Nozzle / Source) Transition point
During the troubleshooting phase, it was ascertained that there must be a more
complex fault with mutual influences between the first and second stages besides
the simple fault caused by sticking spool in the second stage (cf. Sect. 1.2.4.1). It
seems to be a more complex fault with a self-induction, whereas the entire system is
simultaneously influenced by numerous nonlinear parameters. In the worst case, the
whole valve system can even be completely ‘frozen’.
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