Hydraulic force regulating system

Abstract: A hydraulic force regulating system (10') presenting a hydraulic actuator (11) including a cylinder (15) housing a piston (16) movable in relation to the cylinder; a hydraulic circuit (12-14) in turn presenting a servovalve (12) for controlling supply to the actuator; a first force regulating loop (26-29) which, on the basis of the difference between the actual force (F m ) generated by the piston (16) and the required nominal force (F r ), regulates the current (I) supplied to the servovalve (12); and a second regulating loop (35) which, on the basis of the speed (V) of the piston (16) in relation to the cylinder (15), generates an additional current (I₁) for the servovalve (12), to achieve a substantially elastic performance of the hydraulic system (10'). More specifically, the additional current (I₁) opens the servovalve (12) by the exact amount required to permit the passage, in the presence of a desired pressure difference Δp, of the flow AV (product of piston area A and speed) produced by the movement of the piston (16) in relation to the cylinder (15).

Description: The present invention relates to a hydraulic force regulating system. Such a system may be applied to advantage to the active lateral suspensions of railroad cars, to which reference is made herein to give a clear idea of the problem underlying the present invention. Between the body (frame) and the truck of a railroad car, lateral suspensions are provided to absorb the shock and lateral stress on the truck, as shown in Figure 1, which shows a schematic view of a railroad car 1 presenting a truck 2, a body 3, and lateral suspensions 4. The suspensions of most currently operated trains substantially comprise a system of fairly elastic springs and dampers for absorbing shock and preventing stress from being transferred to the body. Such a solution, however, fails to operate correctly when cornering, in which case, the centrifugal force produced shifts the body as far as the limit stops where the suspensions are fully deformed and therefore no longer capable of absorbing lateral stress. By way of a solution to the problem, pneumatic active suspensions have been devised, each of which comprises a pneumatic cylinder connected to a supply tank by valves controlled by an electronic regulating unit. The lateral stress on the truck is absorbed by the flexibility of the suspension itself, while the force generated on the body when cornering is compensated by the regulating system. More specifically, an accelerometer measures the centripetal acceleration of one truck on the train (e.g. that of the engine) correlated to the centrifugal force, and an electronic control unit so controls supply to the pneumatic cylinders as to generate a force in opposition to the centrifugal force. Though theoretically solving the problem, in actual practice, the force generated by the active suspension cylinders involves errors, especially under dynamic conditions, so that the predetermined correction fails to provide a sufficient guarantee of comfort. A hydraulic system has therefore been proposed, which, by virtue of its intrinsic characteristics (rapid response and precision), provides for a more accurate cornering acceleration correction, and for more rapidly and accurately recentering the body in relation to the truck. Moreover, as compared with pneumatic systems, hydraulic systems are more suitable for generating large forces, and are more lightweight and compact. Hydraulic systems, however, present a poor degree of flexibility, so that they fail to provide for absorbing rapid stress such as the shock generated on the trucks of railroad cars. The same also applies in general to hydraulic force regulating systems to which general reference is made below. As stated, hydraulic regulating systems are normally used for generating large forces, and when a rapid, precise variation in the regulated force is required. The possibility of safely employing high pressures, in fact, provides for reducing component weight and size, while the low elasticity of the hydraulic fluid provides for a rapid variation in pressure and hence in the regulated force. The latter property (low elasticity of the hydraulic fluid), however, represents a drawback when regulating a force in the presence of rapid movement of the mechanical member on which the force is exerted, which movement results in an undesired variation in pressure and hence in the regulated force. It is an object of the present invention to provide a hydraulic force regulating system designed to fully exploit the advantages of hydraulic regulation, to ensure precise regulation of the force even in the presence of rapid stress, and to overcome the drawbacks typically associated with known systems. According to the present invention, there is provided a force regulating system comprising a hydraulic actuator including a cylinder housing a piston movable in relation to the cylinder to generate an actual force; a hydraulic circuit supplying said hydraulic cylinder and in turn comprising a control element for controlling supply to the cylinder; and a force regulating loop in turn comprising a measuring unit for measuring said actual force, a differential unit receiving an actual force signal from said measuring unit and a nominal force signal, and generating an error signal, and a regulator receiving said error signal and generating a first control signal for controlling said control element; characterized in that it comprises a speed measuring device for measuring the speed of said piston in relation to said cylinder; and compensating means for generating an additional control signal for said control element on the basis of said speed, and in such a manner as to achieve a substantially elastic performance of said hydraulic system. In practice, the present invention provides for a regulation which, on the basis of the measured speed of the mechanical element on which the regulated force is exerted, varies the regulating signals so as to eliminate, at each instant, any variation in force produced by the movement of the mechanical element. In other words, the regulation according to the present invention provides for rendering a conventional hydraulic system "artificially" flexible. A preferred, non-limiting embodiment of the present invention will be described by way of example with reference to the accompanying drawings, in which: Figure 1 shows a diagram of a railroad car; Figure 2 shows the control principle of a known hydraulic force regulating system; Figure 3 shows a more detailed diagram of part of the Figure 2 system; Figure 4 shows a hydraulic force regulating system in accordance with the present invention. The known control principle of a known force regulating system will be described with reference to Figure 2. The hydraulic system 10 in Figure 2 comprises a linear actuator 11 connected by a servovalve 12 to a supply conduit 13 and a return conduit 14. Actuator 11 comprises a cylinder 15, and a piston 16 movable inside cylinder 15 and connected to a translating mechanical member 18 (in the case of the railroad car in Figure 1, translating member 18 may comprise truck 2, and cylinder 15 may be fixed to body 3). A conduit 20 with a calibrated orifice 21 connects the two chambers 15a, 15b of cylinder 15; and servovalve 12 regulates the supply of pressurized fluid to and from chambers 15a, 15b on the basis of the sign and amplitude of input current I. System 10 also comprises two pressure transducers 22a, 22b connected to and for measuring the pressure in respective chambers 15a, 15b, and which are connected by respective leads 23a, 23b to an electronic regulating unit 25 comprising: a first adding node 26 supplied with the two pressure signals p 1 , p 2 generated by transducers 22a, 22b, and generating a difference signal Δp; a gain unit 27 which multiplies difference signal Δp by a gain coefficient A, and generates a signal F m corresponding to the force F exerted by piston 16; a second adding node 28 supplied with actual force signal F m and with a required force signal F r , and generating a difference signal E equal to the error between the two; a regulator 29 supplying the current I corresponding to the required correction in the control of servovalve 12. In system 10, electronic unit 25 receives the required force signal F r (in the example shown of a railroad car, the force required to eliminate the centrifugal force when cornering, and determined, as stated, by means of a processing unit or a table reading) and the signal F m indicating the actual force generated by actuator 11. On the basis of error E, and using a known regulating principle (e.g. a PID regulator), unit 25 generates a control current signal which is amplified to supply current I to servovalve 12. In the Figure 2 system, conduit 20 and calibrated orifice 21 provide for reducing the pressure gain of servovalve 12, which is too high for certain types of application (such as the railroad car in the example shown), and for consequently improving the stability of system 10. System 10 may be provided with other performance-improving components (not shown), e.g. for reducing offset and leakage on the servovalve, and achieving a given performance in the event of a breakdown. As stated, if the mechanical member 18 subjected to the force F generated by hydraulic actuator 11 is stationary or moving at slow speed, system 10 provides for regulating the force accurately and rapidly. For example, for hydraulic systems capable of generating forces of several tons, it is possible to achieve a frequency response with a passband up to about 10 Hz, and hence a predominant time constant of 15-20 ms. Conversely, if mechanical member 18 is moving at high (constant or variable) speed, even serious errors in the value of the force applied may arise, depending on the operating conditions involved. According to the present invention, hydraulic system 10 is made "artificially" flexible, to permit regulation even in the presence of stress due to the movement of translating member 18. To demonstrate the principles underlying the hydraulic system according to the invention, Figure 3 shows a more detailed view of the structure of servovalve 12, and more specifically the structure of slide valve 30 of the servovalve. For the Figure 3 circuit to achieve a force F in the direction shown (rightwards), slide valve 30 must be moved leftwards in relation to the center position, to permit the passage of a certain amount of fluid from supply line 13 to left chamber 15a, and at the same time permit the passage of the same amount of fluid from right chamber 15b to return line 14. When piston 16 is stationary, the flow Q of fluid through servovalve 12 equals the flow Q L through calibrated orifice 21 and which may be considered roughly proportional to pressure difference Δp = p 1 - p 2 : Q L = Δp * K L where K L is the escape coefficient through orifice 21. In a servovalve, the linearized relationship between flow Q and controlled pressure difference Δp is given by: Q = G Q (I - Δp/G p ) where I is the current of the servovalve, G Q the gain in flow, and G p the gain in pressure. When the piston is stationary, if Q equals Q L (continuity equation) and hence (1) equals (2), pressure difference Δp and servovalve current I present the following relationship: Δp = G Q * I (K L + G Q /G p ) Consequently, a variation in the servovalve current results in a corresponding variation in the pressure difference and hence in the regulated force. What is more, the above equation applies regardless of the position of the piston, so that, whatever position the piston is in, a given servovalve current I gives the same pressure difference Δp. This also means that, when the piston is moving, the movement of the piston in itself produces no variation in pressure. As shown below, it is the speed at which the piston is moving that produces a variation in the pressures in chambers 15a, 15b of cylinder 15. When piston 16 moves rightwards in relation to cylinder 15 at speed V, as shown in Figure 3, and disregarding the compressibility of the fluid, flow Q through servovalve 12 must equal the sum of flow Q L through calibrated orifice 21 and flow AV (the product of the speed of the piston multiplied by its area A) produced by the movement of the piston. In this condition, and taking into account equation (1), the continuity equation becomes: Q = Q L + AV = K L * Δp + AV Substituting (2) in (4) and a number of straightforward passages give a pressure difference Δp in this case of: The numerator term AV/G Q constitutes a disturbance which generates an error in the pressure difference and hence in the regulated force; and, since area A of the piston is selected according to the available pressure and the maximum force to be generated, a large flow gain G Q , i.e. a large servovalve, is required to reduce the pressure error produced by the speed of the piston. That is, all other conditions being equal, the larger the servovalve is, the smaller is the error produced by the speed of the piston. For each application, however, there is a limit to the maximum size permissible of the servovalve. An increase in the size of the servovalve, in fact, results in an increase in internal leakage and hence in the continuous flow required; and an increase in flow gain G Q results in an increase in the gain of the regulating loop, which can only be increased so far without jeopardizing the stability of the system. As such, the above force regulating system presents a high impedance, i.e. a high ratio between the force generated F and displacement speed V. To drastically reduce the impedance of the system as so make it flexible to displacement, the effect of term AV/G Q in (5) must be eliminated or at least considerably reduced, to achieve which, the present invention provides for a speed transducer for measuring the speed of piston 16 in relation to cylinder 15, and a stage for generating a correction current I 1 . Figure 4 shows a regulating system 10' which, as compared with the known system 10 in Figure 2, also comprises a further regulating loop 35 within regulating unit 25'. Loop 35 in turn comprises a compensating stage 36 supplied with a speed signal V generated by a transducer 37 associated with piston 16; and stage 36 generates a correction current I 1 which is supplied to an adding node 38 by which it is added to the current I generated by regulator 29 to supply a total regulated current I 2 to servovalve 12. Compensating stage 36 is none other than a multiplier, which receives speed signal V and generates current I 1 according to the equation: I 1 = K * V This therefore gives a regulated pressure difference Δp of: If the correction factor K value is so selected that K=A/G Q , the disturbance produced in the regulated pressure by the speed of piston 16 is completely eliminated, the impedance of hydraulic system 10' is zero, and piston 16 may therefore move at any speed with no variation in the regulated pressure. This is achieved, in practice, by correction current I 1 of servovalve 12 opening it by the exact amount required to allow the passage, in the presence of the required pressure difference Δp, of the flow AV produced by the movement of piston 16 in relation to cylinder 15. The above solution permits several features for improving the degree of accuracy according to the speed of piston 16. For example, since flow gain G Q , as opposed to being constant, varies alongside a variation in the regulated pressure difference, multiplication correction factor K = A/G Q may be constant and selected according to the average value of G Q , or may vary by following the variations in flow gain with p 1 and p 2 (e.g. using a prememorized table, as shown symbolically by arrow 40 in Figure 4). Clearly, changes may be made to the regulating system as described and illustrated herein without, however, departing from the scope of the present appended claims. In particular, the system according to the present invention may be applied to advantage to the active lateral suspension of a railroad car, to ensure recentering when cornering, by generating a force in opposition to the centrifugal force produced, and also to absorb any shock on the truck capable of producing rapid displacement of the piston, by rendering the system "artificially" elastic as described above. The block diagram shown may be modified as required. In particular, the power unit for generating the current supplied to the servovalve may be separated from regulator 29 and compensating unit 37 and located downstream from adding node 38, in which case, regulator 29 and compensating unit 37 generate respective electronic control signals which are added in node 38 and subsequently amplified.

Claims: A hydraulic force regulating system (10') comprising a hydraulic actuator (11) including a cylinder (15) housing a piston (16) movable in relation to the cylinder (15) to generate an actual force; a hydraulic circuit (12-14) supplying said cylinder and in turn comprising a control element (12) for controlling supply to the cylinder; and a force regulating loop in turn comprising a measuring unit (22a, 22b, 26, 27) for measuring said actual force, a differential unit (28) receiving an actual force signal (F m ) from said measuring unit and a nominal force signal (F r ), and generating an error signal (E), and a regulator (29) receiving said error signal and generating a first control signal (I) for controlling said control element (12); characterized in that the force regulating loop comprises a speed measuring device (37) for measuring the speed (V) of said piston (16) in relation to said cylinder (15); and compensating means (36) for generating an additional control signal (I 1 ) for said control element (12) on the basis of said speed, and in such a manner as to achieve a substantially elastic performance of said hydraulic system (10'). A hydraulic system as claimed in Claim 1, characterized in that said compensating means (36) comprise a multiplier for multiplying a speed signal (V), generated by said speed measuring device (37), by a multiplication factor (K). A hydraulic system as claimed in Claim 2, characterized in that said multiplication factor (K) equals A/G Q , where A is the area of said piston (16), and G Q the flow gain of said control element (12). A hydraulic system as claimed in Claim 3, characterized in that said multiplication factor (K) is constant. A hydraulic system as claimed in Claim 3, characterized in that said multiplication factor (K) is variable, and depends on the pressure difference (Δp) in said cylinder (15) of said actuator (11). A hydraulic system as claimed in any one of the foregoing Claims, characterized in that it comprises an adding element (38) receiving said first control signal (I) and said additional control signal (I 1 ), and generating a sum signal (I 2 ) which is supplied to said control element (12). A hydraulic system as claimed in any one of the foregoing Claims, characterized in that said control element comprises a servovalve (12). A hydraulic system as claimed in any one of the foregoing Claims, characterized in that it constitutes an active lateral suspension (4) of a railroad car (1).

Inventors: BALOSSINI GUALTIERO, BOZZOLA PIERO, JACAZIO GIOVANNI, BALOSSINI, GUALTIERO, BOZZOLA, PIERO, JACAZIO, GIOVANNI
Applicants: MICROTECNICA, MICROTECNICA S.P.A.
Application year: 1995
Main classification: B61F5
IPC codes: B61F5