Patent Publication Number: US-2022228637-A1

Title: Dampar and seat suspension mechanism

Description:
TECHNICAL FIELD 
     The present invention relates to a damper and a seat suspension mechanism using the damper. 
     BACKGROUND ART 
     Various dampers (shock absorbers) for absorbing vibration during traveling are provided in a seat suspension mechanism for supporting a seat of an automobile, a suspension disposed between a vehicle wheel and a vehicle body, and the like. Various dampers are also used for vibration absorption, shock buffering, and the like not only in vehicles such as automobiles but also in various industrial apparatus, joint parts of robots, and opening/closing parts and hinge parts of doors, personal computers, and so on. As such dampers, a damper (a viscous damper or an oil damper) having a viscous liquid filled in its cylinder and using viscous resistance generated by the sliding of a piston in the viscous liquid as in Patent Document 1, a friction damper using a frictional force between a piston and a cylinder as in Patent Document 2, and so on are known. 
     PRIOR ART DOCUMENT 
     Patent Document 
     Patent Document 1: Japanese Patent Application Laid-open No. 2015-78725 
     Patent Document 2: Japanese Patent Application Laid-open No. 2015-117754 
     DISCLOSURE OF THE INVENTION 
     Problems to Be Solved by the Invention 
     The use of the viscous resistance of the viscous liquid as in Patent Document 1 may not enable to obtain a sufficient damping force in a small-stroke, low-velocity zone. Further, if the stroke of an input is small, the friction damper of Patent Document 2 may become a rigid body without functioning as a damper because of a large friction damping force. 
     The present invention was made in consideration of the above and has an object to provide a damper that has a simple structure yet is capable of increasing a damping force according to a stroke amount and thus is capable of contributing to an improvement in damping characteristics in a seat suspension mechanism and so on, and to provide a seat suspension mechanism using the damper. 
     Means for Solving the Problems 
     To solve the above problem, a damper of the present invention includes: 
     an outer cylinder; 
     a movable inner cylinder having a shorter axial-direction length than an axial-direction length of an inner space of the outer cylinder and movable in the outer cylinder; 
     a piston around whose outer peripheral surface a linear member that generates a frictional force with the movable inner cylinder is wound, and which operates in the axial direction in the movable inner cylinder; and 
     movable inner cylinder cover members provided at axial-direction ends of the movable inner cylinder, 
     wherein an orifice is formed in a peripheral wall of the movable inner cylinder, and a viscous liquid is filled in the outer cylinder including the inside of the movable inner cylinder and the nonwoven fabric, and 
     wherein, when the piston moves relative to the movable inner cylinder, a predetermined damping force is generatable owing to viscous friction generated between the piston and the movable inner cylinder and owing to viscous resistance of the viscous liquid passing through the orifice, whereas when the piston moves in the outer cylinder together with the movable inner cylinder, the predetermined damping force is not generated. 
     Preferably, a buffer member impregnated with the viscous liquid is provided in a gap between each of the movable inner cylinder cover members and each end wall of the outer cylinder. 
     Preferably, in the movable inner cylinder and the movable inner cylinder cover members, at least inner surfaces are lower in friction coefficient than the viscous liquid. 
     Preferably, the viscous liquid is low-penetration grease. 
     Further, a seat suspension mechanism of the present invention is a seat suspension mechanism which is disposed between a vehicle body structure and a seat, the seat suspension mechanism including: 
     a spring mechanism through which an upper frame as the movable member attached to the seat is elastically supported on a lower frame as the fixed member attached to the vehicle body structure; and 
     the aforesaid damper which exerts a damping force that absorbs energy generated when the upper frame moves up and down relative to the lower frame. 
     Effect of the Invention 
     According to the present invention, the movable inner cylinder slidable in the axial direction in the outer cylinder and having the orifices formed in its peripheral wall is provided, the piston around which the linear member that generates a frictional force is further provided in the movable inner cylinder, and the viscous liquid is filled in the outer cylinder including the inside of the movable inner cylinder. Accordingly, when the piston moves together with the movable inner cylinder, a high damping force is not exerted, whereas when the piston moves relatively in the movable cylinder, the high damping force is exerted owing to the viscous friction generated between the piston and the movable inner cylinder and the viscous resistance of the viscous liquid passing through the orifice. Therefore, the present invention, when applied to a seat suspension mechanism and the like, achieves efficient vibration damping near a balance point owning to the operation of the spring mechanism, while capable of increasing a damping force near up/down movement ends. Further, in the case where the buffer member having a function of absorbing the viscous liquid is provided between each of the movable inner cylinder cover members and each of the end walls of the outer cylinder, the damping force at stroke ends more increases. 
    
    
     
       BRIEF DESCRIPTION OF DRAWINGS 
         FIG. 1  is a sectional view illustrating a damper according to one embodiment of the present invention. 
         FIG. 2  is a Lissajous figure illustrating the damping characteristics of the damper according to the embodiment. 
         FIG. 3  is a schematic mechanism view of a seat suspension mechanism using the damper according to the embodiment. 
         FIG. 4  is a view illustrating an arrangement relation of the damper, a magnetic spring, torsion bars, and so on in the seat suspension mechanism in  FIG. 3 . 
         FIG. 5  is a chart illustrating the load-deflection characteristics of the seat suspension mechanism in  FIG. 3 . 
         FIG. 6  is a chart illustrating the results of a vibration test regarding EM 6 . 
         FIG. 7  is a chart illustrating the results of a vibration test regarding EM 7 . 
         FIG. 8  is a chart illustrating the results of a vibration test regarding EM 8 . 
         FIG. 9  is a chart illustrating the result of a damping test. 
         FIG. 10  are views illustrating a structure in which a pantograph link mechanism is added to the seat suspension mechanism, (a) illustrating a state in which an upper frame is at the uppermost end, (b) illustrating a state in which the upper frame is at a balance point, and (c) illustrating a state in which the upper frame is at the lowest end. 
         FIG. 11  are views illustrating a structure in which a tensile coil spring and a damper for pantograph are removed from  FIG. 10  for explaining the movement of a fixed-side link and a movable-side link of the pantograph link mechanism, (a) illustrating a state in which the upper frame is at the uppermost end, (b) illustrating a state in which the upper frame is at the balance point, and (c) illustrating a state in which the upper frame is at the lowest end. 
         FIG. 12  is a chart illustrating the load-deflection characteristics of the structure in which the pantograph link mechanism is added to the seat suspension mechanism. 
     
    
    
     DESCRIPTION OF EMBODIMENTS 
     The present invention will be hereinafter described in more detail based on embodiments illustrated in the drawings.  FIG. 1  is a view illustrating a damper  1  of this embodiment. The damper  1  of this embodiment is of a telescopic type having an outer cylinder  2  and a movable inner cylinder  3  and further having a piston  5  which operates in the movable inner cylinder  3  in the axial direction, and so on. 
     The outer cylinder  2  is closed with end walls  22 ,  23  provided at ends of its peripheral wall  21 . In an inner space  24  of the outer cylinder  2 , the movable inner cylinder  3  whose axial-direction length is shorter than the axial-direction length of the inner space  24  (the distance between inner surfaces of the end walls  22 ,  23  at the ends) is provided. 
     The movable inner cylinder  3  has orifices  31  a formed in its peripheral wall  31 . In this embodiment, the two orifices  31   a  are formed close to each of axial-direction ends  31   b  ,  31   c  of the peripheral wall  21 . The number of the orifices  31   a  formed and their pore size may be any as long as they can produce a desired viscous resistance, and depend on the viscosity and so on of a viscous liquid. Further, movable inner cylinder cover members  32 ,  33  are attached to the ends  31   b  ,  31   c  to close them. 
     Between the inner surface of the peripheral wall  21  of the outer cylinder  2  and the outer surface of the peripheral wall  31  of the movable inner cylinder  3 , ball members  4  are provided to contribute to a reduction in sliding resistance when the movable inner cylinder  3  relatively moves in the outer cylinder  2 . 
     The piston  5  has a predetermined length in the axial direction and is provided in the movable inner cylinder  3 . A piston rod  51  is connected to the radial-direction center of the piston  5 , and an end portion  51   a  of the piston rod  51  penetrates through the movable inner cylinder cover member  32  of the movable inner cylinder  3  and the end wall  22  of the outer cylinder  2  to project to the outside. An attachment portion  5  lb provided on the end  51   a  of the piston rod  51  and an attachment portion  23   a  provided on the other end wall  23  of the outer cylinder  2  to protrude outward are attached to a control target. 
     Around the outer peripheral surface of the piston  5 , a linear member  52  is wound. The linear member  52  is densely wound with almost no gap left in the axial direction on the outer peripheral surface of the piston  5 . The linear member  52  is formed of a thread or a cord. Its material is not limited and it is formed of synthetic fiber, natural fiber, or the like. Further, the thread can be either a monofilament or a multifilament but is preferably formed of a multifilament composed of a fiber bundle because such a multifilament facilitates the adhesion of a later-described viscous liquid  7  and is capable of coming into contact with the inner surface of the peripheral wall  31  of the movable inner cylinder  3  at multiple points and thus a high friction damping force can be expected from this. In either of the cases, short fibers may be planted on the surface of the thread or cord forming the linear member  52 , or the surface of the thread or cord forming the linear member  52  may be raised. Such planting of the short fibers or raising facilitates the adhesion of the viscous liquid  7 . 
     The wire diameter (outside diameter) of the thread or cord forming the linear member  52  is not limited either but is selected such that a predetermined frictional force can be exerted on the inner surface of the movable inner cylinder  3  at the time of the movement relative to the movable inner cylinder  3 . Further, depending on its wire diameter, it can be wound in multiple layers such as double layers or triple layers around the outer peripheral surface of the piston  5 . 
     The gaps between the movable inner cylinder cover members  32 ,  33  of the movable inner cylinder  3  and the end walls  22 ,  23  of the outer cylinder  2  are each provided with a buffer member  6 . The buffer members  6  are formed of a material having a function of absorbing the viscous liquid  7  filled in the outer cylinder  2 , for example, a nonwoven fabric impregnatable with the viscous liquid  7 . As such a nonwoven fabric, brand name “FELIBENDY (registered trademark)” of KURARAY CO., LTD. is usable, for instance. The natural length of the buffer members  6  is preferably longer than the gaps (ranges indicated by reference signs A, B in  FIG. 1 ) between the movable inner cylinder cover members  32 ,  33  of the movable inner cylinder  3  and the end walls  22 ,  23  of the outer cylinder  2 . Preferably, the buffer members  6  each have a length about equal to or more than the total length of the two gaps (for example, 20 mm or more if the total length of the two gaps A, B is 20 mm). Consequently, the buffer members  6  can exhibit an enhanced shock absorbing function when the movable inner cylinder  3  is at stroke ends. 
     The viscous liquid  7  is filled in the outer cylinder  2  including the inside of the movable inner cylinder  3 . Considering the adhesion to the linear member  52 , the viscous liquid  7  is preferably low-penetration grease. The low-penetration grease mentioned here is preferably one whose JIS K 2220 penetration at 25° C. is within a range of 175 to 340, more preferably within a range of 265 to 295. 
     Especially in the case where such grease is used as the viscous liquid  7  as in this embodiment, at least the inner surfaces of the movable inner cylinder  3  and the movable inner cylinder cover members  32 ,  33  which constitute flow paths of the viscous liquid  7  are preferably lower in friction coefficient than the viscous liquid  7  so that the fluidity of the viscous liquid can be increased. Specifically, as the movable inner cylinder  3 , one made of stainless steel at least whose inner surface is mirror-finished is usable. Further, the movable inner cylinder cover members  32 ,  33  may be those molded from a thermoplastic resin with a potassium titanate fiber added (brand name “POTICON” manufactured by Otsuka Chemical Col., Ltd.). 
       FIG. 2  illustrates the results of the measurement of the damping characteristics of the damper  1  of this embodiment. In the damper  1  used for the measurement, the gaps between the movable inner cylinder cover members  32 ,  33  of the movable inner cylinder  3  and the end walls  22 ,  23  of the outer cylinder  2  (free play zones (the ranges indicated by reference signs A, B in  FIG. 1 )) are each 10 mm and their total length is 20 mm. Excitation waveforms used here were sine waves whose velocity was highest when the displacement was 0 mm, and their frequencies were 1.0 to 5.0 Hz corresponding to the vicinity of a resonant frequency of a seat suspension mechanism for automobiles. 
     As illustrated in  FIG. 2 , the damping force of the damper  1  of this embodiment does not greatly change when the displacement is between about −7 and about +7 mm. This indicates that the movable inner cylinder  3  is displaced in the axial direction together with the piston  5 . That is, this is a change within the free play zones and the damping force does not greatly change. However, since the buffer members  6  impregnated with the viscous liquid  7  are provided in the free play zones, the acceleration due to the movement of the movable inner cylinder  3  is absorbed. As the displacement more increases, the damping force changes to a certain amount or more. This is because the movable inner cylinder cover members  32 ,  33  of the movable inner cylinder  3  are displaced in the free play zones to gradually compress the buffer members  6  and accordingly, when the displacement reaches a predetermined amount or more, the relative displacement of the piston  5  in the movable inner cylinder  3  starts. Consequently, viscous friction acts between the linear member  52  of the piston  5  and the movable inner cylinder  3  and at the same time, the viscous resistance of the viscous liquid  7  passing through the orifices  31   a  acts, resulting in the generation of the predetermined damping force or more. As the movable inner cylinder  3  further approaches the stroke ends, the buffer members  6  are further compressed, so that the damping force increases, which reduces a bottoming feeling. 
     As is understood from the above, the damper  1  of this embodiment not only has the characteristic that its damping force increases according to the stroke amount but also has a characteristic of increasing in the damping force as the excitation frequency is higher and also having velocity dependence. 
       FIG. 3  and  FIG. 4  illustrate a schematic structure of a seat suspension mechanism  1000  to which the damper  1  of this embodiment is applied. This seat suspension mechanism  1000  includes a lower frame  1100  as a fixed member attached to a vehicle structure (not illustrated) and an upper frame  1200  attached to a seat (not illustrated). 
     The lower frame  1100  supports the upper frame  1200  through a parallel link  1300 . 
     The upper frame  1200  is further elastically supported through a spring mechanism including a linear spring  1410  and a magnetic spring  1420 . The linear spring  1410  is composed of torsion bars  1411 ,  1412  provided at portions connected to the upper frame  1200 , in a front link  1310  and a rear link  1320  forming the parallel link  1300 , and bias the upper frame  1200  through the parallel link  1300  in a direction in which the upper frame  1200  becomes apart from the lower frame  1100 . 
     The magnetic spring  1420  is similar to those disclosed in Patent Documents 2 and 3, details of which are omitted, and includes: a stationary magnet unit  1421  fixed to the lower frame  1100  and having a pair of stationary magnets which are provided with, for example, their same poles facing each other; and a movable magnet unit  1422  supported by a frame provided on the upper frame  1200  and having a movable magnet movable up and down between the pair of stationary magnets. When the upper frame  1200  moves up and down relative to the lower frame  1100 , the movable magnet is displaced in the gap between the pair of stationary magnets, and depending on its relative position, the spring characteristic of the magnetic spring  1420  linearly changes. Specifically, if a characteristic that a restoring force increases in the operation direction of an elastic force (restoring force) of the torsion bars  1411 ,  1412  which are linear springs, that is, in such a direction as to separate the upper frame  1200  from the lower frame  1100  is defined as a positive spring characteristic, the magnetic spring  1420  exhibits a negative spring characteristic that the restoring force in the aforesaid direction decreases, in a predetermined displacement range of the load-deflection characteristic. Therefore, in the range where the negative spring characteristic of the magnetic spring  1420  functions, since this negative spring characteristic is combined with a spring constant (positive spring constant) of the positive spring characteristic of the torsion bars  1411 ,  1412 , there is a constant-load zone where a variation of a load value as the whole spring constant is a predetermined amount or less, that is, a zone where the spring constant is substantially zero (preferably, the spring constant is in a range of about −10 N/mm to about 10 N/mm) even if the displacement increases. Therefore, if this range where the spring constant is substantially zero is adjusted to be near a balance point when a person is seated, it is possible to obtain a high vibration absorbing performance. 
     In the seat suspension mechanism  1000 , the damper  1  of this embodiment is obliquely extended between the lower frame  1100  and the upper frame  1200 . Here, the damper  1  is attached obliquely to the lower frame  1100  with an attachment angle of 10 degrees. Further, the seat suspension mechanism  1000  is designed such that the stroke of the upper frame  1200  from the lowest end to the uppermost end is 40 mm. 
       FIG. 5  is a chart illustrating the load-deflection characteristics of the seat suspension mechanism  1000  in the cases of respective load masses. The balance point was set according to the load mass, and a static balance point was set at the A position 32.5 mm apart from the lowest end (0 mm position). However, in a dynamic state, a dynamic balance point was at the B position. A reason why the balance point moves to the B position in the dynamic state is that vibrational energy input from the unsprung part is converted to potential energy of the magnetic spring to generate a lifting force resisting the gravity. That is, when this lifting force exceeds the whole frictional force of the seat suspension mechanism  1000 , the sprung part swings and the dynamic balance point moves. 
     In  FIG. 5 , the dynamic spring constant in the constant load zone is about 4500 N/m including the frictional force of the whole seat suspension mechanism  1000  when the amplitude is ±7.5 mm. At this time, the frictional force was 76 N when the load was light (Weight: 75 kg) and was 114 N when the load was high (Weight: 120 g). Therefore, the dynamic spring constant cannot exceed 4500 N/m. A suitable whole frictional force experimentally found from a free vibration waveform is  80  N for a person weighing 60 kg, and is 133 N for a person weighing 100 kg. The aforesaid values of 76 N at the light-load time and 114 N at the high-load time are close to these experimentally found values. Therefore, phase control is performed from a low-frequency band and an opposite phase is easily formed around the balance point, achieving excellent vibration absorbency. 
     Next, a seat was attached to the upper frame  1200  of the seat suspension mechanism  1000 , and a vibration test was conducted while a subject A was seated on the seat. The subject A is 171 cm tall and weighs 65 kg. 
     For excitation, input spectral classes (ISO 10326-1) stipulated for the types of machines in JIS A 8304:2001 “Earth-moving machinery-laboratory evaluation of operator seat variation” based on ISO 7096:2000 are used, and vibration absorbing performance is determined based on S.E.A.T values (Seat Effective Amplitude Transmissibility factor of the seat) in the respective cases and standard values of a damping test. The input spectral classes of the excitation waveforms are EM 6 , EM 7 , and EM 8 . The input spectral class EM 6  is a standard for “a crawler tractor-dozer with 50,000 kg or less”, and the excitation waveform has a 7.6 Hz dominant frequency and a PSD maximum value of 0.34 (m/s 2 ) 2 /Hz, and a S.E.A.T value of less than 0.7 and a vibration transmissibility of less than 1.5 at a resonance frequency in a vertical axis direction have to be satisfied. The input spectral class EM 7  is a standard for “a compact damper”, and the excitation waveform has a 3.24 Hz dominant frequency and a maximum PSD value of 5.56 (m/s 2 ) 2 /Hz, and a S.E.A.T value of less than 0.6 and a vibration transmissibility of less than 2.0 at a resonance frequency in a vertical axis direction have to be satisfied. The input spectral class EM 8  is a standard for “a compact loader with 4,500 kg weight or less”, and the excitation waveform has a 3.3 Hz dominant frequency and a maximum PSD value of 0.4 (m/s 2 ) 2 /Hz, and a S.E.A.T value of less than 0.8 and a vibration transmissibility of less than 2.0 at a resonance frequency in a vertical axis direction have to be satisfied. 
     Further, the excitation waveform of the damping test is a logarithmic swept sine wave (0.5 to 4.0 Hz) and an input amplitude is a ±8 mm displacement amplitude corresponding to 40% of the total stroke 40 mm of the seat suspension mechanism  1000 . A load mass was a 75 kg rubber weight, which was fixed with a belt. 
     As a vibrator, a 6-axis vibrator of DELTA TOOLING CO., LTD. was used for EM 7 , EM 8 , and the damping test, and a 3-axis electrodynamic vibrator (3-axis vibrator TAS-1000-5 manufactured by IMV Corp., the maximum excitation stroke 60 mm) was used for EM 6 . On each of shake tables of these vibrators, the seat suspension mechanism  1000  to whose upper frame  1200  the seat was attached as described above was set. 
       FIG. 6 ,  FIG. 7 , and  FIG. 8  illustrate the excitation waveforms EM 6 , EM 7 , and EM 8  and their vibration transmissibilities at the time of the vibration test, all of which are test data of the subject A. As is seen from  FIG. 6 , in the case of the excitation waveform EM 6 , a resonance frequency was 1.2 Hz and a gain of a resonance peak was 1.58. The gain can be kept as low as 0.92 though increasing near 2.0 Hz. Further, at 7.6 Hz where the acceleration of the excitation waveform reaches the peak, the gain reduces to 0.22. This is because the resonance frequency and the resonance gain could be set low owing to the damper  1  of this embodiment. Further, the damper  1  and the frictional force of about  100  N contribute to a reduction in a damping factor and it was possible to decrease the gain in a high-frequency band. Note that the S.E.A.T. value was 0.53, which satisfied the aforesaid standard. 
     As is seen from  FIG. 7 , the resonance frequency of EM 7  was 1.2 Hz and a gain of a resonance peak was 0.99. Regarding EM 7  whose vibration amplitude is the largest among the excitation waveforms in ISO 7096:2000 standard and which is difficult to conform, a gain at 3.24 Hz where the acceleration of the excitation waveform reaches the peak decreases to 0.3. 
     This indicates that, because the damper  1  has the characteristic of changing in damping force according to the amplitude of the excitation waveform, the vibration damping mechanism sufficiently worked on the large-amplitude excitation waveform with a low frequency which is near the resonance frequency of the seat suspension mechanism  1000 . Note that the S.E.A.T value was 0.59, which satisfied the aforesaid standard. 
     As is seen from  FIG. 8 , the resonance frequency of EM 8  was 1.4 Hz and a gain of a resonance peak was 1.0. The peak frequency of the acceleration of the EM 8  excitation waveform is 3.3 Hz, which is approximately equal to that of EM 7 , but a gain decrease degree is smaller as compared with the vibration transmissibility of EM 7 . This is because the acceleration of this excitation waveform is low and the damping factor becomes high owing to the damping force of the damper  1  and the frictional force of the whole seat suspension mechanism and accordingly the gain becomes high at high frequencies. Note that the S.E.A.T value was 0.66, which satisfied the aforesaid standard. 
     As is seen from  FIG. 9 , the resonance frequency of the seat suspension mechanism  1000  was 1.05 Hz and a gain of a resonance peak was 1.1. At 1.15 Hz, the gain becomes lower than 1.0 and at 1.7 Hz, the gain becomes lower than 0.5. This indicates that, despite being a single degree of freedom system with a 40 mm stroke, the seat suspension mechanism  1000  has a high vibration damping performance because it has the constant load zone and the damper  1  has the free play zones. Further, it has the characteristic that the dynamic balance point moves from the A position which is the static balance point to the B position as illustrated in  FIG. 5 , has a dynamic spring constant of 4500 N/m, performs the phase control from a low-frequency band, and forms the opposite phase early, and thus is excellent in vibration absorbency. Note that the gain increase at 3.3 Hz is due to the characteristic of the seat. 
     The comparison of  FIG. 6  to  FIG. 9  shows that the resonance frequency and the gain of the resonance peak differ depending on the input condition, and the seat suspension mechanism  1000  has input dependence. 
       FIG. 10  to  FIG. 12  are views illustrating a structure in which a pantograph link mechanism  200  is added between the lower frame  1100  and the upper frame  1200  of the above-described seat suspension mechanism  1000 . As illustrated in  FIGS. 11( a ) to ( c ) , the pantograph link mechanism  200  have: two fixed-side links  210 ,  220  whose ends (fixed member-supported portions)  211 ,  221  are supported by the lower frame  1100  which is a fixed member, through shaft members  2111 ,  2211  and which are arranged to open wider in a substantially V-shape as they go toward the other ends (fixed member-connection portions)  212 ,  222 ; and two movable-side links  230 ,  240  whose ends (movable member-supported portions)  231 ,  241  are supported by the upper frame  1200  which is a movable member, through shaft members  2311 ,  2411  and which are arranged such that a virtual line connecting these form a substantially inverted V-shape from the ends (movable member-supported portions)  231 ,  241  toward the other ends (movable-side connection portions)  232 ,  242  when the upper frame  1200  is at the uppermost end position. 
     The movable-side links  230 ,  240  are substantially triangular-shaped in a side view, and in their portions projecting more upward than virtual lines connecting the ends (movable member-supported portions)  231 ,  241  and the other ends (movable-side connection portions)  232 ,  242 , further have damper engagement portions  233 ,  243  where to hang a damper (hereinafter, a “damper for pantograph”)  400 . 
     As illustrated in  FIGS. 10( a ) to ( c ) , between the other ends (movable-side connection portions)  232 ,  242  of the movable-side links  230 ,  240 , a tensile coil spring  300  is suspended. Further, a point where the movable member-supported portions (shaft members  2311 ,  2411 ) which are the ends  231 ,  241  of the movable-side links  230 ,  240  overlap with a straight line connecting fulcrums of the tensile coil spring  300  (the ends (movable-side connection portions)  232 ,  242 )) in a side view is a dead point P of the movable-side links  230 ,  240 . The movable-side links  230 ,  240  are rotatable in both directions toward the one-side region and the other-side region sandwiching the dead point P. 
     According to this link mechanism  200 , the tensile coil spring  300  biases the movable-side links  230 ,  240  such that their orientations become opposite when the upper frame  1200  is more upward and when it is more downward than the dead point P. Therefore, the tensile coil spring  300  biases the ends (movable member-supported portions)  231 ,  241  of the movable-side links  230 ,  240  upward when the upper frame  1200  is at a position higher than the balance point, and biases the ends (movable member-supported portions)  231 ,  241  of the movable-side links  230 ,  240  downward when the upper frame  1200  is at a position lower than the balance point. 
     Consequently, when the upper frame  1200  is at a position lower than the balance point, the ends (movable member-supported portions)  231 ,  241  of the movable-side links  230 ,  240  are biased downward. At this time, the load of the tensile coil spring  300  is highest at the dead point P, and the load decreases both when the upper frame  1200  goes upward and when it goes downward. 
     If the setting is made such that the movable-side links  230 ,  240  are at the position of the dead point P when the upper frame  1200  of the seat suspension mechanism  1000  is near the balance point, the ends (movable member-supported portions)  231 ,  241  of the movable-side inks  230 ,  240  are biased downward when the upper frame  1200  is displaced to a lower position than the dead point P, so that a negative spring characteristic of the tensile coil spring  300  is exhibited. 
     Therefore, in the structure including the pantograph link mechanism  200 , not only the negative spring characteristic of the magnetic spring  1420  acts but also the negative spring characteristic of the tensile coil spring  300  is superimposed. This makes it possible to form the constant load zone even if those having a higher spring constant are used as the torsion bars  1411 ,  1412  which exhibit the positive spring characteristic, enabling to cope with a further increase of a load mass. Further, when the upper frame  1200  is displaced from the position lower than the dead point P to a position higher than the dead point P, the negative spring characteristic of the tensile coil spring  300  of the link mechanism  200  tries to restrain this movement, also enabling an improvement in the shock absorbing function. 
     The damper  400  for pantograph is suspended between the damper engagement portions  233 ,  243  as described above. The kind of the damper  400  for pantograph usable here is not limited, but it is preferable to use a damper, disclosed in WO2018/025992, in which a viscous liquid such as grease is stuck to the periphery of a piston sliding in a cylinder and which can exhibit a high damping force as a whole owing to the combination of a viscous frictional force with a viscous damping force. It is also possible to adopt one having the same structure as that of the above-described damper  1  which is obliquely suspended between the upper frame  1200  and the lower frame  1100  of the seat suspension mechanism  1000  and has the free play zones. In this case, the damping force is small near the balance point, and near the end points of the operating range, that is, near the uppermost end and the lowest end of the upper frame  1200 , the largest damping force is exhibited. Therefore, owing to the spring characteristic that the spring constant becomes substantially zero near the balance point, vibration is absorbed, and for a large-amplitude input, owing to the operation of the damper  400  for pantograph in addition to the aforesaid operation of the damper  1 , it is possible to alleviate an impact force and increase the shock absorbing function of reducing the feeling of touching the bottom and touching the top, near the uppermost end and near the lowest end. 
     The seat suspension mechanism  1000  and the pantograph link mechanism  200  are arranged in parallel, and the pantograph link mechanism  200  acts to restrain the movement of the seat suspension mechanism  1000 . Therefore, phases where their operations become effective deviate from each other, achieving the efficient operation against vibration and an impact force. 
       FIG. 12  illustrates the measurement results of the load-deflection characteristics of the seat suspension mechanism  1000  that employs the pantograph link mechanism  200 . Note that one disclosed in WO2018/025992 was used as the damper  400  for pantograph. As illustrated in this drawing, it is seen that an about 15 mm constant load zone is formed near the balance point. Further, the dynamic spring constant in the drawing is 31472 N/m at the maximum because the spring force of the tensile coil spring  300  suspended to the pantograph link mechanism  200  produces the frictional force between the shafts and the frictional force becomes 374 to 403 N. However, the dynamic spring constant decreases as the seat suspension mechanism  1000  moves. 
     EXPLANATION OF REFERENCE SIGNS 
       1  damper 
       2  outer cylinder 
       22 ,  23  end wall 
       3  movable inner cylinder 
       31  peripheral wall 
       31   a  orifice 
       32 ,  33  movable inner cylinder cover member 
       4  ball member 
       5  piston 
       51  piston rod 
       52  linear member 
       6  buffer member 
       7  viscous liquid 
       1000  seat suspension mechanism 
       1100  lower frame 
       1200  upper frame 
       200  pantograph link mechanism