[0001] The present invention falls in the sector of fuel supply systems in internal combustion
engines. More in particular, the present invention refers to a piezoelectric actuator
for the operation of an injection pump of internal combustion engines and to the injector-pump
assembly employing said actuator.
[0002] Supply systems are widely known, in particular for the supply of fuel injectors;
they generally use controlled alternative pumps, also called vibrating pumps, which
are based on the use of an actuator in the shape of a piston of ferromagnetic material
controlled by an electromagnet supplied with alternate current. One of these pumps
is described for example in patent MARELLI
EP-0.953.764, which concerns, however, the oil supply in a two-stroke engine. Another example
is described in
US 6 079 636.
[0003] The present invention proposes to identify an injection system alternative to the
ones already proposed, based on an actuator of a piezoelectric type, and such as to
allow increased injectable flow rates per cycle, over known systems, given equal dimensions
of the actuator and an improved reaction to high operation rpm. It is here reminded
that three pump-injector systems are compared in the known art: with an actuator acting
on a centrally-loaded membrane, or acting on an annular-load membrane, or piston pump-actuator
systems.
[0004] It is evident from the results of the studies carried out on these systems that,
in terms of deliverable fuel flow rate per stroke, in ideal conditions (perfectly
uncompressible fluid, absence of blow-by, infinitely rigid container, ideal valves),
the volume of injectable fluid, in the case of a piston pump-injector, may be obtained
through the following mathematical expression:
wherein ΔV
max = injectable fuel volume per cycle;
ΔL
0 = idle displacement of the actuator;
Δp = difference between injection pressure and supply pressure
F
zbf = load which may be developed by the actuator at the maximum voltage, locked between
two non-yielding restraints (Zero Blocking Force)
[0005] By applying formula (1) to the case of piezoelectric actuators available on the market
with a reference pressure of 75 bar, very modest volumes are obtained, as can be seen
from the table of fig. 1, wherein:
- L
- is the length of the piezoelectric actuator,
- W
- is the actuator diameter
- Fzbf
- is the theoretical load which can be developed by the actuator,
- ΔL0
- is the idle actuator displacement,
whereas in the last column the displaced volumes are shown according to the type of
piston actuator taken into consideration; it is immediately apparent that these volumes
are insufficient, in the majority of cases, to supply even a 50 cc engine. An exception
is only the actuator called EPCOS a which, in the embodiment showing two specimens
arranged stacked (line 2 of the table of fig. 1), is sufficient for the power of a
motorbicycle.
[0006] Due to obvious reasons of greater efficiency, but also for ease of description, here
and in the following reference is always made to a piezoelectric-type actuator applied
to a piston actuator, but it is intended that the teaching of the invention may be
applied also to the other systems mentioned above.
[0007] The object of the present invention is hence to obtain an improvement of the performance
of a piston pump-injector system with piezoelectric actuator, essentially by adopting
means capable of amplifying the useful run of the piezoelectric actuator.
[0008] Such object is achieved by means of a structure of the piezoelectric actuator as
defined in claim 1), as well as by an injection pump structure as defined in claim
4).
[0009] Further features and advantages of the invention will in any case be more evident
from the following detailed description of a preferred embodiment, given purely by
way of non-limiting example and shown in the accompanying drawings, wherein:
fig. 1 is a table showing, as already mentioned above, the data of some commercial
types of piezoelectric actuators;
fig. 2 is a diagram showing the mechanical behaviour feature of a piezoelectric actuator;
fig. 3 diagrammatically shows, in an axial section, a piston pump-injector assembly
with piezoelectric actuator according to the present invention;
fig. 4 shows in an extremely enlarged axial section a Belleville washer used in the
device of fig. 3; and
fig. 5 shows a diagram of the operation feature of a Belleville washer, as it is used
according to the invention.
[0010] As known, and as is evident from the diagram of fig. 2, a piezoelectric actuator
has a linearly decreasing load/deformation feature. That is to say, at a set supply
voltage (phantom lines for 0 V., for 80 V. and for 160 V., respectively), this type
of actuator is capable of developing large loads with small piston displacements,
but loses its thrust capability as the run increases, down to zero thrust when the
maximum idle displacement is reached.
[0011] The diagram of fig. 2 is very easily interpreted: on the y axis the points idle operation
are found, which represent the ideal case set forth earlier, wherein deformation is
due only to the applied voltage (the maximum deformation value corresponding to the
maximum supply voltage is commonly reported as the idle run of the actuator
ΔLo).
[0012] On the x axis, all the zero-deformation points are found; they represent an extreme
situation referring to the non-ideal case, i.e. when the deformation imparted by the
electrical control is fully neutralised by the elastic deformation. At the extreme
point of the y axis, in correspondence of the maximum voltage, it is possible to guess
that the corresponding electrical deformation (for the case set out in the drawing)
at 60 µm at 160 V be neutralised by the elastic deformation equal to F/k, where F
is the load and k is actuator rigidity.
[0013] As can again be guessed from the diagram of fig. 2, to the load of 1140 N a deformation
of 60 µm, i.e. a rigidity of 19 N/µm, must correspond.
[0014] In intermediate situations between idle and zero-deformation, a progressive improvement
of one performance is accomplished at the expense of the other, according to the linear
law of fig. 2.
[0015] As can be easily guessed, the actual feature of the piezoelectric actuator does not
perfectly suit the application thereof as a hydraulic pumping element, because the
latter one requires the exertion of a constant force throughout the entire compression
run. Instead, it has not been possible yet to use the full nominal run of the piezoelectric
element, because, as seen with respect to the diagram of fig. 2, in the final part
of the run the loading capability necessary to overcome fluid pressure would not be
provided.
[0016] Since the piezoelectric actuator cannot exert a force which remains constant throughout
the entire work run ΔL
o, it is provided, according to a first aspect of the present invention, to adopt a
compromise between run and load, making use of a part only of the actual total work
run. In other words, it is provided to use the piezoelectric element with an average
run and an average load. It is possible to prove that this is the best solution, i.e.
the one which allows to obtain the maximum work per cycle; as a matter of fact, in
such solution the largest quantity of energy is delivered between the one available
per work cycle.
[0017] The piezoelectric material of which commercial actuators are made further has the
feature - common to all ceramic materials - of displaying an asymmetric mechanical
behaviour in the presence of tensile and compressive stresses; in particular, tensile
behaviour is poor, whereas the compressive one is acceptable for applications as actuator,
both in static and in dynamic conditions.
[0018] In order to have a duration and reliability compatible with automotive industry requirements,
it is hence imperative, in the light of what has been set forth above, that the actuators
employed always work under a compressive stress. According to another aspect of the
present invention, it is therefore provided to achieve this result by applying a preload
to the piezoelectric element. Such technical practice allows to remarkably increase
the useful life of the actuator.
[0019] From a functional point of view, the presence of a preload (assumed to be rigidly
constant) does not affect the performance of the actuator, determining only a translation
of the work cycle in the plane of fig. 2, i.e. bringing the work cycle inside the
positive half-plane (compressive stress) of the load; thereby, a safety margin is
accomplished which serves to safeguard the piezoelectric element from the dynamic
tensile actions which are generated during fast operation and which must be neutralised
precisely by the presence of the preload.
[0020] Should the preload be accomplished through a contrast spring, which is the simplest
and cheapest solution, it must be noted, however, that when a contrast spring is added,
the performance of a pump actuated by a piezoelectric actuator worsens as contrast
spring rigidity increases.
[0021] According to an essential aspect of the present invention, however, it is suggested
to adopt a negative-rigidity spring, i.e. a spring supplying a decreasing preload
as deformation increases. An actuator deficiency is thereby countered by the special
feature of the spring, i.e. the capability of providing high loads only at the beginning
of the run.
[0022] In the case of the EPCOS actuator identified in line 3 of the table of fig. 1, if
it is intended to work with a fixed preload of 500 N, with an idle displacement of
60 µm and a (fixed) load of 1140 N, only 50% of the run can be exploited. However,
if, according to the present invention, a spring having a rigidity for example of
-9,5 N/µm is used, which is capable of giving an initial-run load of 900 N and a final-run
load of
it was possible to ascertain that the load loss which may be developed by the actuator
is compensated by the spring feature, and 100% of the run can be exploited, thereby
doubling performance.
[0023] Among the springs widespread in the technical practice, the ones capable of displaying
a negative rigidity are conical disc springs, or so-called Belleville washers (see
fig. 4). In this kind of springs, upon varying of the ratio η between the free height
h of the conical disc and its thickness t, different curve trends are obtained; the
diagram of fig. 5 shows an example, wherein:
- if η < 1.41, the load increases in the deformation direction,
- if <1.41 η < 2.83, there is a deformation area, around the free height value, where
the load decreases (negative rigidity);
- if η > 2.83 the previous behaviour is heightened and even a load sign inversion is
obtained (the spring pulls instead of pushing and, if it is not constrained, jumps
into a stable position with higher deformation values).
[0024] The most suitable behaviour for the application of the present invention is evidently
the one where 1.41 < η < 2,83.
[0025] However, the embodiment comprising springs having a negative rigidity of -9.5 N/µm,
although theoretically possible, is of no practical usefulness. As a matter of fact,
in order to obtain a similar value of the diagram gradient of fig. 2, albeit limited
to the middle portion where the derivative takes up the absolute minimum value, it
would be necessary to resort to diameters of 800 mm with thicknesses of 20 mm. These
dimensions are evidently not admissible for use in internal combustion engines, in
particular those intended for motor bicycles. The problem has therefore arisen of
how to manufacture a spring featuring maximum negative rigidity with the smallest
dimensions.
[0026] According to another important aspect of the present invention, it was therefore
resorted to the idea of artificially amplifying the negative rigidity of the spring,
arranging two springs in series and precisely one negative-rigidity spring coupled
with a positive-rigidity spring. It is known that the overall rigidity of two springs
arranged in series is given by the formula:
(which among other things is confirmed by the fact that, when two conventional springs
are placed in series, rigidity decreases: for example, assuming that k
1 = 100 and k
2 = 100, the result is k = 50). If, however, according to the present invention, a
negative-rigidity spring is arranged in series with a positive-rigidity spring, as
mentioned, an overall negative rigidity is obtained, but amplified in its modulus.
For example, assuming k
1 = 100 and k
2 = -95, the result becomes k = 1900).
[0027] On the basis of such a configuration, it is hence possible to design a spring capable
of providing maximum negative rigidity with respect to material resistance, having
dimensions compatible with its application in internal combustion engines. As a result
of this dimensioning, a spring having the following measures is for example manufactured
(see the references of fig. 4) :
D = 28 mm
d = 16 mm
h = 1.3 mm
t = 0.5 mm
h/t = 2.6
k
min = -0.56 N/µm
phosphate-containing steel material 1.1248 C75S(Ck75)
wherein a high ratio between the spring height
h and the thickness
t of the spring disc is highlighted.
[0028] The rigidity thus obtained for the spring dimensioned as above is then doubled with
two springs arranged packetwise and further amplified arranging in series a conventional
spring with a rigidity of 1,15 N/µm.
[0029] From the calculations carried out, it appeared that a piezoelectric pump-injector,
particularly suited to small-powered internal combustion engines - having for example
a swept volume of 50 cm
3 - can be manufactured, according to the present invention, also employing a small,
inexpensive piezoelectric element such as the Epcos one (table of fig. 1), in association
with a preload spring having a negative rigidity whose modulus is about half that
of the piezoelectric element.
[0030] It is intended, however, that the invention is not to be considered limited to the
particular arrangement illustrated above, which represents only an exemplary embodiment
thereof, but that several changes are possible, all within the reach of a person skilled
in the field, without departing from the scope of protection of the invention, as
defined in the following claims.
1. Piezoelektrische Betätigungseinrichtung für die Betätigung einer Einspritzpumpe für
Verbrennungskraftmotoren, umfassend in Kombination ein piezoelektrisches Element und
eine Gegendruckfeder, die aus zumindest einer Belleville-Scheibe besteht, wobei die
genannte Gegendruckfeder zumindest eine Belleville-Scheibe mit negativer Steifigkeit
aufweist, d.h. die eine mit zunehmender Verformung abnehmende Kraft bereitstellt,
und die so bemessen ist, dass das Verhältnis zwischen der Federhöhe h und der Dicke
t der Belleville-Scheibenfeder in dem Bereich 1.41 < h/t < 2.83 liegt, und dadurch gekennzeichnet, dass zumindest eine herkömmliche, eine positive Steifigkeit aufweisende Feder in Reihe
mit der genannten Belleville-Scheibe angeordnet ist, wobei die genannte Gegendruckfeder
eine Vorspannung des piezoelektrischen Elements liefert.
2. Piezoelektrische Betätigungseinrichtung nach Anspruch 1, dadurch gekennzeichnet, dass die genannte Belleville-Scheibe eine negative Steifigkeit aufweist, deren Modul etwa
die Hälfte der Steifigkeit des piezoelektrischen Elements beträgt.
3. Piezoelektrische Betätigungseinrichtung nach Anspruch 1, dadurch gekennzeichnet, dass sie ein Paket von drei identischen Belleville-Scheiben mit negativer Steifigkeit
umfasst, die parallel zueinander und in Reihe mit zumindest einer herkömmlichen, eine
positive Steifigkeit aufweisenden Feder angeordnet sind.
4. Einspritzpumpe für Verbrennungskraftmaschinen, insbesondere mit kleiner Leistung,
dadurch gekennzeichnet, dass sie eine Betätigungseinrichtung nach einem der vorangehenden Ansprüche aufweist,
in Kombination mit einer Kolbenpumpen-Einspritzvorrichtung.