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  • 专利说明
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    • 专利摘要:
    The invention relates to a system for controlling the relative displacement of a load P comprising at least one main longitudinal displacement damping means (1) with displacement C, and having two ends, one being connected to a frame and the other being related to the load, characterized in that it comprises a compensation device comprising at least one secondary damping means of longitudinal action having two ends, one of which is integral with said frame and the other is connected to the end of the main damping means related to the load, and in that said secondary damping means is arranged so that at a point of the stroke C, the secondary damping means has a steering action orthogonal to the direction of said displacement.
    公开号:FR3025788A1
    申请号:FR1552883
    申请日:2015-04-03
    公开日:2016-03-18
    发明作者:Du Chaffaut Benoit Amaudric
    申请人:IFP Energies Nouvelles IFPEN；
    IPC主号:
    专利说明:

    [0001] The present invention is part of the field of mechanical bonds between two elements, or sets of elements, whose relative positions vary between limits determined by the connecting member, and in which it is desired to control the connection force between the two elements to make it practically independent of the relative positions of said elements. In particular, it is desired to keep the binding force substantially constant so that the present invention can be applied in the field of lifting, or load damping. The invention comprises a particular connection between two objects or mechanical assemblies, one being called "support", and the other "load". The purpose of the support is to limit the movement of the load when it is subjected to various stresses, such as its own weight. The connections usually encountered between a support and a load are either rigid or deformable. The deformable bonds are most often elastic, that is to say that the connecting force depends on the elongation of the connecting member, and are then called suspensions. For example, when an automobile (load) moves with its wheels (support) on uneven ground, the suspension means, deforming, allow the wheels to follow the irregularities of the ground without the car is too shaken. However, the deformation of the dampers causes a variation of the suspension force, and the automobile, whose mass remains a priori unchanged, is therefore subject to more or less comfortable vertical accelerations. However, these variations in suspension forces remain indispensable by the elastic return function they provide, an unavoidable function in terms of suspension where the load can vary very substantially due to dynamic phenomena. An area where suspension problems are a little different is that of floating marine installations. Indeed, as soon as the depth of water is too great to place the installation on the bottom, the support, vessel or platform, is necessarily floating and is therefore subject to various uncontrollable movements, including vertical heave due to the swell. A load can be simply suspended from the float, and accompany it in its movements, the only variations in the suspension force 3025788 2 then result eventually from its inertia or the resistance opposed by water to its vertical movements. The load can also be more or less related to a fixed point at the bottom of the sea. The suspension effort must then be as constant as possible whatever the position of the float, under penalty of costly damage to the most fragile elements of the sea. load. Two types of solutions are then possible: 1. Either the operation is limited in time, and the load is not too important. This is the case for example of laying a load of a few tons at the bottom of the sea from a ship subjected to a strong swell. The most effective solution is a dynamic servo of the hoisting system that provides, or dissipates, the energy necessary to maintain the load substantially stationary when the support moves. This is called "active compensation", and the load displacement performance of the load can be centimeter when the displacements of the support are metric. 2. Either the load is very high, permanent or has a long operating time.
    [0002] The cost of energy required for active compensation becomes prohibitive. This is the case for vertical risers connecting an underwater well head to a floating platform for a period of 10 to 20 years, or a set of rods partially laid at the bottom of the well in the process of drilling by a drill ship for several days or weeks. The conventional solution, called "passive compensation", consists most often of hydropneumatic cylinders, ie air springs. These systems do not consume energy, allow large vertical displacements (several meters), have almost linear characteristics (stress-elongation curves), and allow a great variety of situations depending on the volumes and the pressure of the reservoirs. air. Performance is measured by varying the suspension effort as a function of displacement, and strongly depends on the available quantities of high pressure air. Thus, to compensate for a train of stems of about 300 tons being drilled on a ship subjected to a swell of 7 to 8 meters in amplitude (25 ft.), That is to say to limit the variation of the weight on the tool to plus or minus 2% of the total hanging weight, or about 6 tons, it is necessary to have a reserve of air at 200 bars of more than 40 m3.
    [0003] An improved passive compensation system for offshore drilling operations has been described in US5520369 (FR 2575452), where the necessary air volume mentioned in the example above has been reduced by about half, thanks to the implementation of a particular geometry of the cable way of the lifting system 5 providing a complement to the force of the hydropneumatic cylinders. In practice, ships and platforms assigned to underwater work are very often equipped with both systems, and use them in a complementary and simultaneous way if necessary. However, the limited capabilities and operating costs of the active systems still encourage more performance and accuracy for the passive systems. The object of the present invention is to further improve this concept of passive compensation, and possibly to extend its field of application.
    [0004] Thus, the present invention relates to a system for controlling the relative displacement of a load P comprising at least one main damping means of longitudinal action with displacement C, and having two ends, one being connected to a built and the other being tied to the load. It comprises a compensation device comprising at least one secondary longitudinal damping means having two ends, one of which is integral with said frame and the other is connected to the end of the main damping means connected to the load, and said secondary damping means is arranged such that at a point of the stroke C, the secondary damping means has a directional action orthogonal to the direction of said displacement. The longitudinal action damping means may be of the spring cylinder, hydraulic, or pneumatic type, or a combination. The stroke C may correspond at most to the length of the rod of said cylinder. At least two secondary damping means may be arranged symmetrically with respect to the axis of said main damping means, so that their actions cancel out when they are orthogonal to the axis of displacement.
    [0005] The orthogonality point may be substantially in the middle of the stroke C. One of the ends of the secondary longitudinal damping means may be connected to the end of the main damping means related to the load. intermediate of an articulated system.
    [0006] Advantageously, the articulated system comprises a connecting rod. According to one embodiment, the articulated system comprises a first connecting rod whose first end is fixed to the end of the main damping means connected to the load, and a second connecting rod having a first end articulated with respect to a second end of the first link and a second end hinged to the frame, the articulation between the first link and the second link being fixed at the end of the secondary damping means connected to the main damping means. Alternatively, the articulated system comprises a first connecting rod whose first end is fixed to the end of the main damping means connected to the load, and a second connecting rod having a first end articulated with respect to a second end of the first connecting rod. , a hinge relative to the frame, and a second end attached to the end of the secondary damping means connected to the main damping means.
    [0007] The longitudinal displacement of the secondary damping means may be in a fixed direction relative to the direction of movement of the main damping means. The invention also relates to a heave compensator of a floating support, which comprises a load displacement control system as described above. In the compensator, the main damping means may consist of at least two steering cylinders substantially parallel to the direction of the load. The main and secondary damping means may consist of 25 hydraulic cylinders. The primary and secondary damping means may comprise independent hydropneumatic means for regulating their hydraulic pressure independently.
    [0008] The present invention will be better understood and its advantages will appear more clearly on reading the description which follows, of exemplary embodiments, in no way limiting, and illustrated by the appended figures, and among which: FIGS. la-d schematically represent a so-called "isodyne" compensator with T-shaped compression springs, - FIGS. 2a-d schematically represent an "isodyne" compensator with combined springs and connecting rods. FIG. 3 shows an exemplary response of FIG. an isodyne compensator with 3 T-springs, - FIG. 4 shows an exemplary response of an isodyne compensator with springs and connecting rods, - FIGS. 5a-c show schematically an isodyne compensator with hydropneumatic cylinders 10 in three positions, - FIG. 6 shows an exemplary response of a hydropneumatic isodyne compensator with 3 T-jacks, - FIGS. 7a-c show an isodyne compensator diagram. combined cylinders and connecting rods; FIG. 8 shows the theoretical behavior of a pneumatic isodyne compensator with 3 T-cylinders for different load cases; FIG. 9 shows the theoretical behavior of a pneumatic isodynomulator with combined cylinders and rods. FIG. 10 shows the theoretical behavior of a pneumatic isodyne compensator with jacks and connecting rods for different load cases; FIGS. 1a, 1b and 1c show schematically the application of a hydropneumatic isodyne compensator to a lifting system on floating support. FIGS. 12a, 12b, 12c and 12d schematically represent a so-called "isodyne" compensator with an articulated system comprising two connecting rods. FIGS. 13a, 13b, 13c and 13d show schematically an alternative embodiment of the so-called "isodyne" compensator with an articulated system comprising two connecting rods. FIG. 14 illustrates a geometrical parameterization of an articulated system comprising two connecting rods. FIGS. 15a to 15e show a heave compensator comprising an "isodyne" compensator according to the embodiment variant of FIGS. 13. FIG. 16 shows the theoretical behavior of an isodyne compensator with an articulated system comprising two connecting rods according to variant FIG. 17 represents an isobar compensator comprising an isodyne compensator according to the invention. - Figure 18 shows a suspension system comprising an isodyne compensator according to the invention. FIG. 19 represents curves for different stiffnesses, of the stroke as a function of the lift for a suspension equipped with an "isodyne" compensator. According to the invention, the force produced by an elastic spring is substantially proportional to its deformation (arrow). It therefore varies more or less linearly between two extreme values (minimum and maximum). The desired constant value (setpoint) lies between these extremes, for example in the middle. To maintain the same set point value throughout the race (maximum - minimum), it is necessary to add or subtract from the main spring force a complementary force, substantially proportional to the deflection of the deflection. relative to the position corresponding to the reference force, here for example at the midpoint. A first solution proposed by the present invention consists in arranging one or more lateral springs, with intersecting axes at one of the ends of the main spring, and perpendicular to the axis of the same main spring for the reference value. The principle of such compensation will be better understood by means of the example described in FIG. 1. The four representations 1a, 1b, 1c and 1d show the device according to four positions of the moving point M. FIG. set point, where no resultant force of the auxiliary or secondary, lateral springs acts on the point M. The main spring exerts the force Fpo. In the remainder of the description and for all the embodiments described, the term "spring" designates a damping means which may take the form of a spring cylinder, a hydraulic cylinder, a pneumatic cylinder or a hydraulic cylinder. an analogous system or a combination of such cylinders. The main compression spring 1 provides the force Fp ',' when it is compressed to its minimum length Lp ,,,, (Fig. La), and the force Fpn ,,, when it is relaxed to its maximum length Lp ',' (Figure ld). At the half of the stroke C, the effort is worth the reference value Fpo (figure lb). Figure lc illustrates the operation at any point of the race, marked by the value x, this effort is evaluated by the formula: Fp = Fpo - Kpx 3025788 7 where Kp is the stiffness of the main spring. Two compression springs 2, 3, identical to each other, called secondary, are arranged symmetrically on either side of the axis of the main spring. They have a common movable end hinged to the movable end M of the main spring 1, and their fixed ends 4, 5 are articulated on the same bearing structure as that of the main spring. These springs are aligned on an axis perpendicular to that of the main spring when the latter is halfway (Figure lb) at a distance B. For all other positions, the vertical component of the resultant of their efforts is added at 10 the force of the main spring for the upper half of the race, and retrench for the lower half. With similar notations to those of the main spring, one can then write: Fs = Fsmax - Ks (Ls Lsmin) with: Ls - Lsmin = V B2 + B The resulting effort P, or lift of the system, is written: P = Fp + 2Fssin oc with: sin oc = Let finally: 2x P = F -K x + p0 px2 + B2 [F smax K ( 132 + x, 2 B) 1 Isodyne compensation (ie ie constant effort) then amounts to choosing the quantities and parameters allowing, if not checking the equation below, at least 20 to minimize the first term over the largest possible part of the race. 2x ix, 2 + B2r [Fsmax -Ks (a2 + .x2 -B)] - Kpx = 0 In practice, the characteristics of the springs are imposed. However, it is still possible to add extensions or fittings of adaptable lengths, or combine them in series or in parallel. The variable x is limited by the travel of the main spring, and the magnitude B is decisive for the inclination of the secondary springs.
    [0009] One method of determining the set of parameters may be optimization by means of a computer model that makes it possible to explore a large number of combinations by quickly visualizing the result.
    [0010] It will not be departing from the scope of the present invention to change the number or arrangement of the secondary springs, for example by taking up the horizontal component of their force by mechanical guidance.
    [0011] A second solution, fairly close mechanically, consists of interposing articulated rods between the movable end M of the main spring and secondary secondary springs Si and S2 (Figure 2a). It is then necessary, for the equilibrium of the forces, to take again the efforts of the legs of the rods by a suitable guidance, perpendicular to the axis of the main spring. FIGS. 2a-d illustrate this solution in which the secondary or auxiliary springs do not incline with respect to the orthogonal to the axis of the main spring, but compress and still relax in the same axis thanks to the action of the rods 6, 7 of length Lb, and appropriate guides. The forces of the springs are written: Fp = Fp0 - Kpx Fs = Fsmax Ks (Ls Lsmin) with: L s - L smins = Lb -, 1 L2b - X2 The equilibrium of a connecting rod implies that the horizontal components Fbh and vertical Fb, of its compression Fb are linked by: FbV = Fbht9OC with Lb length of the connecting rod: tg a = L2b - x2 and at any point of the race: Fhh Fs and: P = Fp + 2Fb, 20 L ' expression of the lift P is then written: 2x P = F p0 - Kpx + x2 F smax -xv -KS (Lb -, 114 - x2) 1 The compensation "isodyne" amounts to checking the equation: 3025788 9 [ Fsmax - Ks (Lb - _ 1L2 1L2b - x2 b - over the largest possible part of the race For nonzero x, this last equation can be simplified to: 2x Kpx = 0 Kp Fsmax - KsLb Ks 2 + = 0 - '/ L2b -x2 It therefore appears that there is at least one "perfect" compensation solution, at least in the static domain: it concerns any set of quantities and parameters such that: K = n and L b = s 2 Fsmax Ks that will provide isod compensation yne theoretically ideal on the entire race. In this implementation, it will not depart from the scope of the invention by changing the number or arrangement of secondary springs, for example by replacing two opposing compression springs by a tension spring positioned between the two legs of the 10 connecting rods, and producing the same efforts. A third solution, quite mechanically close to the second solution, consists in inserting a system articulated between the mobile end M of the main spring and those of the secondary springs S1 and S2 (FIG. 12a). According to the embodiment illustrated in Figures 12a to 12d, each articulated system comprises two rods 8 and 9 and 10 and 11, respectively, hinged relative to each other. The first connecting rod 9 (or respectively 11) has an end fixed to the end of the main spring 1 connected to the load and a second end articulated with respect to a first end of the second connecting rod 8 (or 10). The second rod 8 (or 10) has an articulated end 20 relative to the second end of the first rod, and a second articulated end 13 (or 15) relative to the frame. The end of the secondary spring 2 (or 3) connected to the main spring is attached to the hinge 12 (or 14) between the first rod 9 (or 11) and the second rod 8 (or 10). The main springs 1 and 2 of this solution are arranged and mounted relative to the frame in a manner identical to their arrangement and mounting for the first and second solutions. FIGS. 12a-d illustrate this solution in which the secondary or auxiliary springs do not tilt with respect to the orthogonal to the axis of the main spring, but compress and still relax in the same axis thanks to the action of rods 8, 9, 10 and 11 of length Lb, and appropriate guides.
    [0012] For this third solution, the system of articulated rods is lighter and easier to build. The trajectories of the connecting rod feet are circular arcs and the points of application of the forces exerted by the secondary springs can be moved, or used as a lever to reduce the bulk or lower the center of gravity of the assembly. In this implementation, it will not depart from the scope of the invention by changing the number or arrangement of secondary springs, for example by replacing two opposing compression springs by a tension spring positioned between the two legs of the rods, and producing the same efforts.
    [0013] A fourth solution, quite mechanically close to the third solution, consists in inserting a system articulated between the moving end M of the main spring and those of the secondary springs S1 and S2 (FIG. 13a). According to the embodiment illustrated in Figures 13a to 13d, each articulated system comprises two rods 8 and 9 and 15 respectively 10 and 11, hinged with respect to each other. The first rod 9 (or 11) has an end fixed to the end of the main spring 1 connected to the load and a second end articulated relative to a first end of the second rod 8 (or 10). The second rod 8 (or 10) has an end articulated with respect to the second end of the first link, a hinge 13 (or 15) relative to the frame, and a second end fixed to the secondary spring 2 (or 3). The hinge 13 (or 15) does not correspond to one end of the rod 8 (or 10) and is disposed between the two ends. The main springs 1 and 2 of this solution are arranged and mounted relative to the frame identically to their arrangement and their assembly for the first and second solutions. FIGS. 13a-d illustrate this solution in which the secondary (or auxiliary) springs do not tilt with respect to the orthogonal to the axis of the main spring, but compress and still relax in the same axis thanks to the action of rods 8, 9, 10 and 11 of length Lb, and appropriate guides. For this fourth solution, the system of articulated rods is lighter and easier to build. The trajectories of the connecting rods are circular arcs and the points of application of the forces exerted by the secondary springs can be moved, or used as levers to reduce the size or lower the center of gravity of the assembly.
    [0014] In this implementation, it will not depart from the scope of the invention by changing the number or arrangement of secondary springs, for example by replacing two opposing compression springs by a tension spring positioned between the two legs of the connecting rods. , and producing the same efforts. For example, it is conceivable to connect two legs of symmetrical rods with a transverse traction spring producing the desired compression in the first rods 9 and 11. The ends of such a tension spring can be found anywhere on the first and second links, provided they are symmetrical about a vertical axis. In another example, the secondary compression springs can be replaced by a single spring connecting the second ends of the second links. According to another embodiment, the linear secondary springs can be replaced by springs or torsion or bending bars controlling the rotation of the rods around the joints around the frame. For the third and fourth solutions, the computation of the forces is carried out as for the first two solutions and according to the geometrical parameterization of FIG. 14. FIG. 14 corresponds to the solution of FIGS. 13a-13d. The force of the main spring is: Fp = Fp0 - Kpx where Fp0 is the nominal force halfway, and Kp the stiffness. The horizontal component Fsh of the force FS produced by the secondary spring at point B provides, at the articulation point 12 or 14 of the two connecting rods, an opposite horizontal force, of intensity Fsh * -n with n the length between the joints the second connecting rod 8 or 10 m with the frame and the secondary spring and m the length between the joints of the second connecting rod 8 or 10 with the frame and the first connecting rod 9 or 11. The end of the main spring connected to the load being limited to a vertical line, this effort induces compression of the BC rod. According to the position of the point C, and therefore according to the inclination of the rod BC relative to the horizontal, this complementary force is directed upwards or downwards, or is zero when the connecting rod BC is horizontal. The computation of the complementary effort for each position of C is based on the resolution of the triangles ABC and APT respectively, by means of the generalized Pythagoras theorem, or cosine law, giving the values of the angles cp, a and 0 as well as the Length PT allowing to know the effort F. By positing beforehand: h = -s2 - g + x and AC2 = h2 + e2, one can write the following equations: 3025788 12 (AC2 + m2 - 12) cf) = arccos 2. AC .m + arctg Eh) O = arcsine 1 PT2 = n2 + AP2 - 2. n. AP .cos (n - - cf) - (3) with = arctg (AP2 + PT2 - AT2) and finally: Fs = Fsmax Ks (PT - PTmin) where: - Fsmax smax is the maximum compression of the secondary spring, corresponding to the minimum PTmin of the length PT, that is to say when the connecting rod BC is horizontal. - Ks is the stiffness of the secondary spring. The vertical complementary effort is therefore: n FS '= FS.-m. Finally, the total lift of the system is obtained by adding the force of the main spring to the complementary force calculated in this way: P = Fp + F 'Examples: 15 1: Compensator with 3 springs in T: By arranging springs trade according to the diagram of the figures la, it is possible to model the behavior of the device. The (compression) springs chosen for this example have the following characteristics: (h - msirup) a = - arccos 2. AP PT 3025788 13 Main spring: Middle spring Olma T2 56 90 500 Load daN Arrow Length mm mm 1000 250 Lj 250 Pc 800 Fc 200 Lc 300 Pb 630 Fb 160 Lb 340 Pa 500 Fa 125 The 375 0 0 LO 500 Lateral springs Light spring Olma T2bis 56 84 400 Load daN Arrow Length mm mm 560 240 Ld 160 Pc 450 Fc 190 Lc 210 Ph 360 Fb 150 Lb 250 Pa 280 Fa 120 The 280 0 0 LO 400 The indices a, b, and c characterize the values respectively at 50%, 63%, and 80% of the stroke of each spring. The distance B of FIG. 1a is the sum of the minimum length of a secondary spring and a value E representing the ends and ball joints (here respectively 177 and 76 mm, ie 253 mm for B). For a nominal load (Fpo, set point) of the main spring of 500 daN, corresponding to half of its arrow, the graph of Figure 3 shows the behavior of the device. The curve CS represents the setpoint, the curve RPS the response with a single main spring, and the curve RC the response with the compensation according to the figures la-d. Over a wide central range of the stroke C, about 100 mm, the lift F is almost constant at about 500 daN. At the ends of the race, the system has a small stiffness providing a minimum elastic return of about 100 daN. An application such as an antivibration filter, or even anti-seismic, could be interesting. 2: Spring and spring compensator (Figures 2a-d): 3025788 14 Using the same main spring as for the example above, and lighter lateral springs associated with articulated rods length 176 mm according to the configuration of the FIGS. 2a-d show the following behavior: lateral springs Light spring Olma T2bis 45 68 250 load arrow mm length mm daN 355 150 Ld 100 Pc 280 Fc 120 Lc 130 Pb 225 Fb 96 Lb 154 Pa 180 Fa 76 La 174 0 0 LO 250 where an improved effect can be seen in FIG. 4, since the constant load path F of 500 daN has been increased from 100 (FIG. 3) to 150 mm. "Ideal" compensation, on the other hand, is more difficult to achieve from the characteristics of the catalogs alone. With springs "custom", it will be possible to approach it more easily, provided in particular that the stiffness values remain constant throughout the race and they do not evolve over time. 3: Compensator with pneumatic or hydropneumatic cylinders (FIGS. 5a-c): The springs on the catalog, or even on measure, have the disadvantage of being limited to a single and rather narrow loading range. Pneumatic or hydropneumatic cylinders, on the other hand, have the advantage of being able to adapt to a load by a simple adjustment of the pressure, and to obtain a stiffness by varying the ratio between the variation in volume in the jack and the volume. total gas to which the system is connected. As generally compressed air, which is not quite a perfect gas, the generally accepted formula for connecting the pressure P and the volume V is: P.VY = constant With for the air: y = 1.4 The response of a pneumatic cylinder can not therefore be linear like that of a coil spring. Its "stiffness" varies slightly along the race.
    [0015] By replacing the springs of FIGS. 1a-d by such jacks, the diagram of FIGS. 5a-c is obtained. The secondary or auxiliary cylinders are connected to a hydraulic circuit P2 V2, the main cylinder being connected to a second Pi circuit V1. Thus, the settings of the hydraulic pressure are independent.
    [0016] 5 The example is based on an application of compensation for heave due to swell. The following table shows the main dimensions of the installation: PCBDNS dns V1 V2 P1 max (bars) P2 max (bars) (t) (m) (m) (m) (m2) (m) (m2) ( m3) (m3) 454 7.62 3.81 0.400 2 0.25132 0.08 4 0.02011 22.3 5.0 191 175 The maximum load is 454 tonnes (1 million pounds: 1000 Klbs or Kips), and 10 the maximum stroke of 7.62 meters (25 ft .) corresponds to a standard length for this type of cylinder. The main cylinders are 2 (N) and the secondary 4 (n). The graph in Figure 6 shows the behavior of the compensator. The curve CS represents the setpoint, the curve VVS the response with a single main spring, and the curve VC the response with the compensation according to FIGS. 5a-c. The entire stroke C of 7.62 m 15 is traveled with a nearly constant lift to 454 tons, the absolute value of the maximum difference being 2,623 tons, or 0.58% of the set value CS. By varying the pressures of each circuit, it is possible to compensate for heave for any lower load case with the same material, as summarized in the table below and the graph in Figure 8: F tonnes 45 90 135 180 225 270 315 360 405 454 V1 n3 22.3 22.3 23 22.3 22.3 22.3 22.3 22.3 22.3 22.3 V2 m3 5.0 5.0 5.0 5.0 5.0 5.0 5.0 5.0 5.0 P1 bars 18.98 38.0 56,) 75.9 94.9 113.9 132.9 151.8 170.8 191.5 P2 bars 17.31 34.6 51.9 69.2 86.5 103.8 121.1 138.4 155.7 174.6 maximum error T 0.260 0.520 0.780 1.040 1.300 1.560 1.820 2.080 2.340 2.623 maximum error% 0.58% 0.58% 0.58% 0.58% 0.58% 0.58% 0.58% 0.58% 0.58% 0.58% 4: Compensator with pneumatic cylinders or hydropneumatic with connecting rods (FIGS. 7a-c): By replacing the springs of FIGS. 2a-d by hydro-pneumatic cylinders, the diagrams of FIGS. 7a-c are obtained. The load case and the race are identical to the previous example. The table below shows the main dimensions: 20 3025788 16 PC Lb DNS dns V1 V2 P1 ma, (bars) P2, '' (bars) (t) (m) (m) (m) (m2) (m) (m2) (m3) (m3) 454 7.62 5.497 0.386 2 0.23404 0.10 4 0.03142 14.906 0.197 209.8 161.9 The diameter of the main cylinder rods has been reduced by 4 mm, and the air volumes have decreased significantly from 22.3 to 14.9 m3 for V1, and 5.0 to 0.197 m3 for V2. The graph in Figure 9 shows the behavior of the compensator. The curve CS 5 represents the setpoint, the curve VVS the response with only one main cylinder, and the curve VC the response with the compensation according to FIGS. 7a-c. The maximum difference is 1,272 tonnes, which is 0.28% of the value of the deposit. The theoretically ideal compensation is very close, but it is in any case illusory because of the non-linearity of the pneumatic response.
    [0017] By carrying out the same scanning of the load cases as for the preceding example, the following summary table and the graph of FIG. 10 are obtained: F tonnes 45 90 135 180 225 270 315 360 405 454 VI m3 14.906 14.906 14.906 14.906 14.906 14.906 14.906 14.906 14.906 14.906 14.906 V2 m3 0.197 0.197 0.197 0.197 0.197 0.197 0.197 0.197 0.197 0.197 Pl bars 20.80 41.6 62.4 83.2 104.0 124.8 145.6 166.4 187.2 209.8 P2 bars 16.05 32.1 48.2 64.2 80.3 96.3 112.4 128.4 144.5 161.9 maximum error T 0.126 0.252 0.378 0.504 0.630 0.756 0.883 1.009 1.135 1.272 maximum error% 0.28% 0.28% 0.28% 0.28% 0.28% 0.28% 0.28% 0.28% 0.28% 0.28% Figures 11a, 11b and 11c illustrate an application of the device according to the invention to a system load lifting device, for example a drilling rig supported by a floating installation. The various operations, drilling (application of a weight on the constant tool), tension on the riser, handling of the wellhead, require control of the displacement of the load P irrespective of the displacement of the floating support at will of the swell. The incorporation of the device according to the invention in the heave compensator system allows an optimization of the control. Figure 1 lb shows the position at the midpoint of the race, where the secondary cylinders connected to the muffle "fixed" (crown block) by connecting rods 6, 7 do not act in addition to the main cylinders. Figure 1 shows the position of the compensator system in the up position with the addition of the forces provided by the secondary cylinders, through the connecting rods. Figure 11c shows the system in the low position with the subtraction of the forces provided by the secondary cylinders, through the connecting rods. The movable block to which the load P is suspended substantially constant, remains substantially immobile with respect to the bottom given the preservation of the length of the cable, thanks to the cable path as the prior art. In FIG. 11a, it will be noted that each of the main and secondary cylinder systems has independent static pressure control means (P2V2 and P1V1). Of course, as described above, the use of rods is not systematic, but greatly facilitates the incorporation of a system according to the invention in a conventional compensator, for example that described in US5520369. 5: Compensator with pneumatic or hydropneumatic cylinders with connecting rods (FIGS. 15a-e): The application of this fourth solution to conventional offshore drilling vessels or platforms, that is to say with cable winch and hydropneumatic cylinders acting on the "fixed" muffle at the masthead, can be carried out simply. Indeed, the articulated bar system of the heave compensator is already used to keep the length of the cable along the race, for example in a conventional heave compensator, including that described in US5520369. Figures 15a to 15e show the principle of such an embodiment for a run of 25 ft. (7.62 m) (only the hydropneumatic circuits are not to scale, and Figures 15a to 15e represent one half of heave compensator, the other half is deduced by symmetry). Figures 15a to 15e illustrate the movement of the main cylinder 1 and secondary 2 over the entire stroke C of the upper muffle compensator heave. The calculation of the lift of such a system is similar to the calculation with mechanical springs detailed above, replacing the stiffness Kp and KS by their pneumatic equivalents calculated at each point by the formula "polytropic" P.VY, constant. An optimization of the conventional case 454 tons x 7.62 m (1000 Kpounds x 25 ft.) Results in the theoretical result detailed in Table 1 and the graph in Figure 16.
    [0018] Table 1 - Example Data 5 Fo s (m) 1 (m) min (m) e (m) g (m) D (m) y (m) z f 3 (°) AP (t) (m) (m) ) (m) 454 7.62 5.60 2.50 1.005 3.931 1.563 0.41 2.2 0.9 22.2 2.38 3025788 18 NV P SP (m2) d (m) NV A SA (m2) V1 (m3) V2 (m3) P1 max (bars) P2 max (bars) and air NP TC (t) 2 0.264 0.235 4 0.173 6 0.4 209.9 195.3 1.4 7 32.4 with the notations of figures 14 and 15 and D the diameter of the rods of the main cylinders, NVP the number of main cylinders, SP the total section of the 5 main cylinders, d the diameter of the rods of the auxiliary cylinders, NVA the number of auxiliary cylinders, SA the total section of the auxiliary cylinders, NP the number of pulleys of the moving block, and TC the tension of the rope. The graph in Figure 16 shows the behavior of the compensator. The curve CS, in dashed line, represents the setpoint, the curve VVS the response with only the 10 main vertical cylinders (V1), the curve LVV the linear response (theoretical) of the vertical main cylinders and the curve VC, in solid line, the response with compensation according to Figures 15a-e. The maximum absolute deviation from the hanging weight setpoint is 1.384 tonnes, or 0.305% of the load. The device is therefore efficient when compared to the prior art where the best results are rather of the order of 2% or more. The main interest of the device according to this embodiment also lies in reducing the total volume of air at high pressure required. The best previous embodiments, require 15 to 20 m3 of high pressure air (210 bar), while the present example is satisfied with 6 m3 of main volume (V1) and 400 liters for the auxiliary circuit 20 (V2), with pressures of the same order. In addition, it can be noted that the lift of the main cylinders only halfway is slightly lower than the set value. In practice, this means that the BC rods are not horizontal at the exact mid-race. The adjustment of the maximum pressure Pl, performed at the zero of the stroke of the main cylinders, is not affected. On the other hand, the maximum of the pressure P2, obtained by putting the connecting rods BC horizontally, is a little higher than the value at halfway (in this case about 0.5 bar), considered as the maximum to start up. 'optimization. This detail must be taken into account in the design. Applications As mentioned for the various examples illustrated above, the system for controlling the relative displacement of a load (according to any one of the embodiments of the invention described above), also called compensator "isodyne". Can be integrated into a heave compensation system used on a floating support. For example, the isodyne compensator can be integrated into a conventional compensator as described in US5520369. Two examples of integration into a heave compensator are shown in Figures 1a-1c and 15a-15e. The use of the isodyne compensator according to the invention in a heave compensator allows an optimization of the passive control of the displacement of the load.
    [0019] For a heave compensator, the frame corresponds to the floating installation. For this application, the main damping means may consist of at least one, preferably at least two, steering cylinders substantially parallel to the direction of the load (essentially vertical). In addition, in the heave compensator, the primary and secondary damping means may comprise hydraulic cylinders. In addition, the main and secondary damping means may comprise independent hydropneumatic means for regulating their hydraulic pressure independently. 2. The displacement control system according to the invention (isodyne compensator) may also be applied to an isobaric pressure compensator. Indeed, any enclosed enclosure (envelope, balloon, pipe, circuit, ...) dimensionally stable containing one or more fluids capable of volume variations (expansion or thermal shrinkage, chemical reaction, ...) must be protected against variations in induced pressure, which are all the stronger as the fluid is not very compressible.
    [0020] This protection consists in most cases in placing the enclosure in communication with a volume of gas, which is considerably more compressible than the fluid in question and is generally isolated from it by a membrane or a piston, and thus limits the amplitude of the the variation of pressure to contain said pressure within the limits allowed by the enclosure. The device, known as a pressure accumulator, is limited by the volume of the gas, which itself constitutes a second chamber subjected to the same constraints as the main enclosure, and by the fact that this volume must be all the more important that we want to limit the amplitude of the pressure variation. A well-dimensioned pressure accumulator constitutes a protection against the destruction of the enclosure, and makes it possible to keep the fluid inside, but its bulk and the regulatory obligations related to the pressure (periodic checks, etc.) in limit the use, especially in cases of expensive or dangerous fluids. However this protection is not absolute in case of uncontrollable expansion (runaway of a chemical reaction, fire, ...).
    [0021] Another type of protection generally used is to open the enclosure on the outside by means of a rupture disk or a valve, as soon as the limit pressure is reached. Such equipment is an absolute protection against the destruction of the enclosure, but the evacuated fluid is lost, and in general the function provided by the enclosure must at least be restarted (case of a chemical reactor). This type of protection is therefore preferable for inexpensive fluids (water, air, etc.) and for cases where the temporary stopping before restarting is not too penalizing. The application of the invention to the protection of closed enclosures containing one or more fluids capable of volume variations can be carried out rather simply according to the diagram of FIG. 17. A cylinder of section A is connected to an enclosure 16 15 filled with a fluid at the pressure P. A sealed piston 17 sliding in the cylinder over a length separates the fluid inside the enclosure from the outside space. An isodyne compensator according to the invention connects said piston 17 to a fixed frame relative to the chamber 16. The isodyne compensator can be formed according to the third solution illustrated in FIG. 12, that is to say with a main spring 1, a secondary spring 2, the secondary spring 2 being connected to the main spring 1 by means of an articulated system comprising two rods 8 and 9. Thus the force exerted on the piston by the isodyne compensator being almost constant, the pressure P reigning inside the enclosure will be very little affected by variations in the volume of the internal fluid causing the movement of the piston in one direction or the other. Such a compensator may further be provided with a piston position sensor for detecting any variation in the volume of the inner fluid. Such a detection makes it possible, for example, without changing the pressure conditions, to better control the evolution of a chemical reaction inside the enclosure, or to show a loss or a gain of fluid with a very small flow rate. . Finally, this device can also be used as a pulsation damper, for example on a piston pump, or as a temporary accumulator of mechanical energy. For this last use, the stored energy is the product of the displacement of the piston by the constant force delivered by the equipment. The advantage is to have the same force throughout the stroke of the piston, which allows for example 3025788 21 to operate a hydraulic motor with an oil pressure independent of the level of "filling" of the accumulator. In the same way, if the same (reversible) hydraulic motor is used as a pump, the filling is also done at constant pressure. The equipment optimized for a single pressure thus has a maximum efficiency over the entire range of use, and the instantaneous power, in storage and drawing, theoretically depends only on the possible oil flow. One conceivable application is the recovery of the braking energy of a vehicle, where the available instantaneous power capacity is greater than the total amount of energy itself. 3. The load displacement control system according to the invention (the "isodyne" compensator) may be further used in a suspension system of a vehicle. Indeed, a mass M suspended on a support by a deformable link (suspension) reacts to the pulses or oscillations of the support according to two main characteristics of the suspension: the stiffness K which determines the force exerted as a function of the amplitude of the deformation, and the damping coefficient C which determines an additional effort depending on the speed of the deformation. M and K define the eigenfrequency, or resonant frequency, of the system, which is written as: f = -27r The essential function of a suspension being to filter the oscillations of the support, a minimum value for K is chosen, so that the natural frequency is as low as possible at the frequencies expected for said oscillations. However, K must be sufficient to prevent the suspension from abutting, for example in the case of significant dynamic forces added. As it is generally not possible to modify K 25 (coil spring, rubber block, ...), the chosen value is a compromise between divergent stresses, and the adaptation of the majority of the suspensions to the different conditions of loading of the support (frequency, dynamic effects, ...) is realized only by playing on the damping coefficient, which is relatively easy to modulate by using hydraulic circuits where a fluid is forced to pass through openings of 30 variable sections. The principle of the invention allows quite easily to continuously adjust the desired stiffness for the entire stroke (travel) of a suspension. Fig. 18 illustrates a vehicle suspension comprising an isodyne compensator according to the fourth solution illustrated in Figs. 13a-13d. The isodyne compensator is disposed between the wheel 19 and the frame of the vehicle 18, for example the chassis of the vehicle. With reference to FIG. 19, it can be seen that it is possible to modulate the compression of the connecting rod 9, and therefore the overall elastic response of the suspension, by simply varying the length of the lever of the rod 8 from zero (no compensation) to a maximum value (isodyne compensation). An equivalent result can be obtained by modifying the lever arm and / or the stroke and therefore the range of possible forces of the secondary spring 2, for example by means of the adjustable length Lr of the connecting rod 8.
    [0022] FIG. 19 illustrates several curves representing the variation of the stroke C as a function of the lift F in tonnes, for different stiffness K. Each stiffness K is associated with a value of the length of the lever n (see FIG. These curves correspond to a suspended mass of 0.5 tonnes and a travel of 500 mm. According to this example, the stiffness of such a device can be adjusted between 2 daN / mm (main spring alone) and 0.001 15 daN / mm, ie about 1 gram / mm (isodyne suspension). Note that only the curve corresponding to the maximum stiffness is a straight line. The slight distortion of the other curves is due to the geometry of the articulated bar system which can not approach the equilibrium of the initial solution with transverse guides of the legs of connecting rods. Other adjustments are possible, in particular the distance separating the pivot A from the minimum of the race of the main spring. The effect then relates to the value of the lift around which we can vary the stiffness. It is possible, up to certain limits, to adapt the suspension to a variation of approximately ± 30% of the median value. 25
    权利要求:
    Claims (16)
    [0001]
    1) A system for controlling the relative displacement of a load P comprising at least one main damping means (1) of longitudinal travel C travel displacement, and having two ends, one being connected to a frame and the other being related to the load, characterized in that it comprises a compensation device comprising at least one secondary damping means (3,
    [0002]
    2) longitudinal action having two ends, one of which is integral with said frame and the other is connected to the end of the main damping means related to the load, and in that, said means of secondary damping is arranged so that at a point of the stroke C, the secondary damping means has a directional action orthogonal to the direction of said displacement. 2) System according to claim 1, wherein said longitudinal action damping means are of the type spring cylinder, hydraulic, or pneumatic, or a combination.
    [0003]
    3) System according to claim 2 wherein the stroke C corresponds at most to the length of the rod of said cylinder.
    [0004]
    4) System according to one of the preceding claims, wherein at least two secondary damping means are arranged symmetrically with respect to the axis of said main damping means, so that their actions are canceled when they are orthogonal to the axis of displacement.
    [0005]
    5) System according to one of the preceding claims, wherein said orthogonality point is substantially in the middle of the stroke C.
    [0006]
    6) System according to one of the preceding claims, wherein one end of the secondary damping means of longitudinal action is connected to the end of the main damping means related to the load via an articulated system. 3025788 24
    [0007]
    7) System according to claim 6, wherein the articulated system comprises a connecting rod (6, 7).
    [0008]
    The system of claim 6, wherein the articulated system comprises a first link (9, 11) having a first end attached to the end of the main load-related damping means (1), and a second connecting rod (8, 10) having a first end articulated with respect to a second end of the first connecting rod (9, 11) and a second end hinged to the frame, the articulation between the first connecting rod (9, 11) and the second connecting rod (8, 10) being attached to the end of the secondary damping means (2, 3) connected to the main damping means (1).
    [0009]
    9) System according to claim 6, wherein the articulated system comprises a first link (9, 11) whose one end is fixed to the end of the main damping means (1) connected to the load, and a second connecting rod (8,
    [0010]
    10) having a first end articulated with respect to a second end of the first connecting rod (9,
    [0011]
    11), a hinge (13, 15) relative to the frame, and a second end attached to the end of the secondary damping means (2, 3) connected to the main damping means (1). 10) System according to one of claims 6 to 9, wherein the longitudinal displacement of the secondary damping means is in a fixed direction relative to the direction of movement of the main damping means. 11) Compensator for heaving a floating support, characterized in that it comprises a load displacement control system according to one of the preceding claims.
    [0012]
    12) Compensator according to claim 11, wherein said main damping means comprises at least two steering cylinders substantially parallel to the direction of the load.
    [0013]
    13) Compensator according to one of claims 11 or 12, wherein the main and secondary damping means consist of hydraulic cylinders. 3025788 25
    [0014]
    14) Compensator according to claim 13, wherein the main and secondary damping means comprise independent hydropneumatic means for adjusting their hydraulic pressure independently.
    [0015]
    15) Isobaric expansion compensator for an enclosure (16) comprising a volume of fluid, the compensator comprising a piston (17) in a cylinder of the enclosure (16), characterized in that the piston (17) is connected to the frame by a system according to one of claims 1 to 10.
    [0016]
    16) Suspension for a land vehicle comprising at least one system according to one of claims 1 to 10 connecting a wheel (19) of the vehicle and the frame (18) of the vehicle, the system comprising a means of adjusting the stiffness of the system. displacement control. 5 10 15
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    同族专利:
    公开号 | 公开日
    FR3025787A1|2016-03-18|
    CN106715315B|2019-03-08|
    FR3025788B1|2019-04-05|
    BR112017003815A2|2017-12-05|
    CN106715315A|2017-05-24|
    KR20170058925A|2017-05-29|
    WO2016041671A1|2016-03-24|
    EP3194320B1|2019-10-23|
    US20170254384A1|2017-09-07|
    US10359093B2|2019-07-23|
    FR3025787B1|2019-06-07|
    BR112017003815B1|2021-08-31|
    EP3194320A1|2017-07-26|
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    法律状态:
    2016-03-18| PLSC| Publication of the preliminary search report|Effective date: 20160318 |
    2016-04-21| PLFP| Fee payment|Year of fee payment: 2 |
    2017-04-26| PLFP| Fee payment|Year of fee payment: 3 |
    2018-04-13| PLFP| Fee payment|Year of fee payment: 4 |
    2019-04-25| PLFP| Fee payment|Year of fee payment: 5 |
    2020-04-29| PLFP| Fee payment|Year of fee payment: 6 |
    2021-04-27| PLFP| Fee payment|Year of fee payment: 7 |
    优先权:
    申请号 | 申请日 | 专利标题
    FR1458698|2014-09-16|
    FR1458698A|FR3025787B1|2014-09-16|2014-09-16|SYSTEM FOR MONITORING THE MOVEMENT OF A LOAD|
    FR1552883A|FR3025788B1|2014-09-16|2015-04-03|SYSTEM FOR MONITORING THE MOVEMENT OF A LOAD|FR1552883A| FR3025788B1|2014-09-16|2015-04-03|SYSTEM FOR MONITORING THE MOVEMENT OF A LOAD|
    EP15750649.4A| EP3194320B1|2014-09-16|2015-07-27|System for controlling the movement of a load|
    US15/511,538| US10359093B2|2014-09-16|2015-07-27|System for controlling the movement of a load|
    KR1020177007046A| KR20170058925A|2014-09-16|2015-07-27|System for controlling the movement of a load|
    CN201580049752.XA| CN106715315B|2014-09-16|2015-07-27|System for controlling load movement|
    BR112017003815-3A| BR112017003815B1|2014-09-16|2015-07-27|CARGO DISPLACEMENT CONTROL SYSTEM, FLOATING SUPPORT COMPENSATOR AND ISOBARIC EXPANSION AND SUSPENSION FOR LAND VEHICLE|
    PCT/EP2015/067141| WO2016041671A1|2014-09-16|2015-07-27|System for controlling the movement of a load|
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