Abstract:
A damper includes a piston rod, a damping piston, at least one cylinder containing a damping liquid, a fixed partition member for partitioning the interior of the damper into two liquid chambers, a pressure source, and a valve in communication with the pressure source which reacts as a function of the pressure. The valve can also be in communication with additional forces, such as mechanical spring forces, which can be adjustable. The valve can include a pressure intensifier. The valve generates fluid flow resistance during flow of liquid in a first direction through the partition member. The fluid flow resistance in the first direction varies according to the amount of force communicated to the valve by the pressure source and any additional forces. The partition member can include means for providing low-resistance return flow of liquid in a second direction.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application is a continuation of U.S. application Ser. No. 11/261,051, filed 27 Oct. 2005, now U.S. Pat. No. 7,374,028 which is a continuation of International Application No. PCT/US2004/038661, filed 18 Nov. 2004, which is a continuation of U.S. application Ser. No. 10/661,334 (now abandoned), filed 12 Sep. 2003, which claims priority to Provisional Application No. 60/485,485, filed 8 Jul. 2003, the entireties of which are hereby incorporated by reference herein and made a part of this specification. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     This invention relates to a damper and, more particularly, to a damper suitably used as a shock absorber or front fork on the suspension of a bicycle, motorcycle, automobile or other vehicle. 
     2. Description of the Related Art 
     Dampers (shock absorbers, MacPherson struts, front forks, etc.) for controlling vehicle body motion and handling characteristics during vehicle travel over uneven surface are well-known in the art. Dampers typically comprise a closed hydraulic cylinder with an internal piston connected to a central piston rod, which reciprocates within the cylinder to produce damping forces. 
     As is well known in the art, the damping forces created by a damper have a major influence on the overall dynamic performance of a vehicle. A wide range of dynamic conditions are encountered during typical vehicle motion over various surfaces and terrain features. For example, these features and conditions include large and small bumps, sharp-edged bumps and round-edged bumps, close-spaced bumps and widespaced bumps, stutter bumps and gradual undulating bumps, and so forth. In addition, conditions include vehicle acceleration and deceleration modes, uphill and downhill travel modes, as well as turning modes. 
     Besides the factors noted above, different operators of a specific vehicle traversing identical terrain features often prefer significantly different damping characteristics. This is especially true for light-weight vehicles, such as bicycles or motorcycles, where rider weight can be a major portion of total weight, and where rider “style” or “technique” can have a significant influence on overall suspension performance. 
     SUMMARY OF THE INVENTION 
     The present invention provides an improved damper which provides automatic modulation of damping forces based on sensing and reacting to internally-generated or externally-generated conditions. 
     In one embodiment, a damper generates a compression damping rate that is modulated in accordance with an internally-generated pressure. An example of an internally-generated pressure is the air or nitrogen pressure found in the wide-variety of conventional “DeCarbon-type” pressurized dampers as have been known in the art for 40 years (reference U.S. Pat. No. 3,101,131 to DeCarbon, issued in 1963). 
     In another embodiment, a damper generates a compression damping rate that is modulated in accordance with an externally-generated pressure. An example of an externally-generated pressure would be the pressure that could be created at an end fitting of a compressed external coil-over spring. 
     In another embodiment, a damper generates a compression damping rate that is modulated in accordance with an independently-regulated pressure. An example of an independently-regulated pressure would be a pressure source controlled by computer and supplied to the shock absorber. The computer may utilize input from various sensors on the vehicle (for example sensors monitoring vehicle speed and acceleration, as well as the relative positions and velocities of the sprung and unsprung masses) and continuously regulate the pressure supplied to the shock absorber in accordance with a pre-determined algorithm. 
     In another embodiment, a damper having damping features may be quickly and easily tuned and adjusted by simply rotating one or more readily-accessible external knobs or levers. Turning an external knob (or knobs) is quick and easy and thus can be done in a routine “on-the-fly” manner frequently during the ride. Since terrain and trail conditions constantly change, this greatly benefits the rider by enabling him/her to continuously select the best damping characteristics for the current situation. 
     In another embodiment, a damper includes valving structures directly adjoining, or within, a fixed partition member in the damper that partitions a portion of the damper interior into two liquid chambers. The valving structures specifically do not directly adjoin, or comprise part of, the main damping piston connected to the piston rod of the damper. The valving structures react as a function of internal or external pressures to provide damping forces by restricting fluid flow in one direction through the fixed partition member. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a sectional front view of a prior-art embodiment of a pressurized damper unit. 
         FIG. 2  is a sectional front view of the prior-art damper of  FIG. 1  modified in accordance with an exemplary embodiment of the present invention. 
         FIG. 3  is an enlarged partial sectional front view of the damper of  FIG. 1 , showing the added structure of this embodiment of the present invention. 
         FIG. 4  is a sectional front view of the prior-art damper of  FIG. 1  modified in accordance with a second exemplary embodiment of the present invention. 
         FIG. 5  is an enlarged partial sectional front view of the damper of  FIG. 4 , showing the added structure of this embodiment of the present invention. 
         FIG. 6  is a sectional view of the damper of  FIG. 5 , taken through section A-A of  FIG. 5 . 
         FIG. 7  is a sectional front view of the damper of  FIG. 4 , showing shaft displacement fluid flow through the fixed partition member during an extension stroke of the damper. 
         FIG. 8  is a sectional front view of the damper of  FIG. 4 , showing shaft displacement fluid flow through the intensifier valve during a compression stroke of the damper. 
         FIG. 9  is a sectional front view of the damper of  FIG. 4  modified in accordance with a third exemplary embodiment of the present invention, with the intensifier valve structure moved to the upper end of the damper cylinder, with a remote reservoir assembly added, and with the floating piston re-located from the damper cylinder to the reservoir cylinder. 
         FIG. 10  is a sectional front view of the prior-art damper of  FIG. 1 , modified in accordance with a fourth exemplary embodiment of the present invention, with the upper eyelet replaced by a piggyback eyelet with an attached reservoir cylinder, with the floating piston re-located from the damper cylinder to the reservoir cylinder, and with the intensifier assembly located in the upper end of the reservoir cylinder. 
         FIG. 11  is an enlarged partial sectional front view of the damper of  FIG. 10 , showing the added structure of this embodiment of the present invention. 
         FIG. 12  is a sectional front view of the damper of  FIG. 10 , showing this embodiment with the addition of an external intensifier adjusting screw. 
         FIG. 13  is an enlarged partial sectional front view of the damper of  FIG. 12 . 
         FIG. 14  is a sectional front view of the damper of  FIG. 10 , with a fifth exemplary embodiment of the present invention located in the upper end of the reservoir cylinder. 
         FIG. 15  is an enlarged partial sectional front view of the damper of  FIG. 14 . 
         FIG. 16  is a sectional front view of the damper of  FIG. 10 , with a sixth exemplary embodiment of the present invention located in the upper end of the reservoir cylinder. 
         FIG. 17  is an enlarged partial sectional front view of the damper of  FIG. 16 . 
         FIG. 18  is a sectional front view of the prior-art damper of  FIG. 1  modified in accordance with a seventh exemplary embodiment of the present invention. 
         FIG. 19  is an enlarged partial sectional front view of the damper of  FIG. 18 , showing the added structure of this embodiment of the present invention. 
         FIG. 20A  is a sectional front view of the prior-art damper of  FIG. 1  modified in accordance with an eighth exemplary embodiment of the present invention. 
         FIG. 20B  is an alternate version of the damper of  FIG. 20A  with modified structure to provide a first alternate shape to the compression damping characteristic produced by the exemplary embodiment of  FIG. 20A . 
         FIG. 20C  is an alternate version of the damper of  FIG. 20A  with modified structure to provide a second alternate shape to the compression damping characteristic produced by the exemplary embodiment of  FIG. 20A . 
         FIG. 21  is an enlarged partial sectional front view of the damper of  FIG. 20A , showing the added structure of this embodiment of the present invention. 
         FIG. 22  is a sectional front view of the prior-art damper of  FIG. 1 , modified in accordance with a ninth exemplary embodiment of the present invention, including elimination of the floating piston. 
         FIG. 23  is an enlarged partial sectional front view of the damper of  FIG. 22 , showing the added structure of this embodiment of the present invention. 
         FIG. 24  is a sectional front view of the prior-art damper of  FIG. 1 , modified in accordance with a tenth exemplary embodiment of the present invention, including elimination of the floating piston and addition of an intensifier preload spring. 
         FIG. 25  is an enlarged partial sectional front view of the damper of  FIG. 24 , showing the added structure of this embodiment of the present invention. 
         FIG. 26  is a sectional view of the damper of  FIG. 25 , taken through section A-A of  FIG. 25 . 
         FIG. 27  is an enlarged partial sectional front view of the damper of  FIG. 24 , modified in accordance with an eleventh exemplary embodiment of the present invention, including elimination of the intensifier preload spring and addition of an intensifier open-bias spring. 
         FIG. 28  is a sectional front view of an air-sprung bicycle shock absorber, modified in accordance with a twelfth exemplary embodiment of the present invention. 
         FIG. 29  is a sectional front view of the prior-art damper of  FIG. 1 , modified in accordance with a thirteenth exemplary embodiment of the present invention. 
         FIG. 30  a sectional front view of the prior-art damper of  FIG. 1  modified in accordance with a fourteenth exemplary embodiment of the present invention. 
         FIG. 31  is an enlarged partial sectional front view of the damper of  FIG. 30 , showing the specific structure of this embodiment of the present invention. 
         FIG. 32  is a sectional front view of a modified version of the damper of  FIG. 30 , incorporating a fifteenth exemplary embodiment of the present invention. 
         FIG. 33  is an enlarged partial sectional front view of the damper of  FIG. 32 , showing the specific structure added to this embodiment of the present invention. 
         FIG. 34  is an overall perspective view of the front fork of a bicycle. 
         FIG. 35  is an overall sectional front view of one leg of the fork of  FIG. 34 , incorporating a sixteenth exemplary embodiment of the present invention. 
         FIG. 36  is a sectional front view of the damper assembly of the fork leg of  FIG. 35 . 
         FIG. 37  is an enlarged partial sectional front view of the damper of  FIG. 36 , showing the specific structure of this embodiment of the present invention. 
     
    
    
     DETAILED DESCRIPTION OF EXEMPLARY EMBODIMENTS 
     The prior-art damper  100  of  FIG. 1  will be described first, in order to provide a point of departure for better understanding the improvements of the present invention, which will be described further on. It is to be understood, of course, that this specific prior-art embodiment is representative only, and that the embodiments disclosed herein may be applied to other types of dampers. 
     In  FIG. 1  the prior-art damper  100 , as is known to those skilled in the art, is comprised of an upper eyelet  110  and a lower eyelet  112  for attachment to, for example, the sprung and un-sprung portions of a vehicle (not shown). The lower eyelet  112  is connected to the piston rod  120  which passes through the seal head  130  and has a damping piston  140  attached at the other end. The damping piston  140  reciprocates in the damper cylinder  150  as the sprung and unsprung portions of the vehicle move relative to each other when, for example, the vehicle traverses uneven terrain. The damping piston  140  has rebound valving  141  (shown symbolically here) and compression valving  142  (also shown symbolically) for restricting fluid flow during rebound strokes (lengthening) and compression strokes (shortening). The valving produces damping forces that resist the imposed motion. For example, the valving structures may be flexible stacks of disc valves covering flow ports through the damping piston  140 , suitable for a variety of applications and condition. 
     Still referring to  FIG. 1 , the damper cylinder  150  is sealed at one end by the seal head  130  and at the other end by the upper eyelet  110 . A floating piston  160  is sealingly engaged, but free to reciprocate, toward the upper end of the damper cylinder  150 . The floating piston  160  separates the hydraulic fluid  170  below it from the internally-pressurized chamber  180  above it, which contains a pressurized gas (for example, nitrogen or air). The Schrader valve  190  provides access to the internally-pressurized chamber  180 , which forms a pressure source. 
     The damping piston  140  divides the total amount of hydraulic fluid  170  contained in the damper cylinder  150  into two portions: a portion above the damping piston  140  (i.e., compression chamber  150   a ), and a portion below it (i.e., rebound chamber  150   b ). When the damping piston  140  moves upward in the damper cylinder  150  (a compression stroke) some of the hydraulic fluid  170  in the compression chamber  150   a  flows downward through the damping piston  140 , via the compression valving  142 , into rebound chamber  150   b . The compression valving  142  restricts this flow, creating compression damping. 
     When the damping piston  140  moves downward in the damper cylinder  150  (a rebound stroke) some of the hydraulic fluid  170  below the damping piston  140  must flow upward through the damping piston  140 , via the rebound valving  141 , into the area above the damping piston  140 . The rebound valving  141  restricts this flow, creating rebound damping. 
     In order to understand the operation of the exemplary embodiments, it is also important to clearly understand the movement of the floating piston  160 , and of the hydraulic fluid  170  below it, during an inward or outward movement of the piston rod  120 . Specifically, it is important to understand the flow of hydraulic fluid  170  that occurs due to the additional volume displaced by the piston rod  120  as it enters the damper cylinder  150 , as well as the flow that occurs due to the volume vacated by the piston rod  120  as it is withdrawn from the damper cylinder  150 . 
     During a compression (upward) stroke such as described above, the piston rod  120  moves further into the damper cylinder  150 , thus occupying more of the total available internal volume of the damper cylinder  150 . The volume occupied by the additional length of the piston rod  120  that enters the damper cylinder  150  displaces an equal volume of the hydraulic fluid  170 , which moves upward and is accommodated by an upward movement of the floating piston  160 . This decreases the volume of the internally-pressurized chamber  180  above the floating piston  160 , which correspondingly increases the pressure somewhat. The net effect is that the added volume of the entering piston rod  120  is accommodated by an equally decreased volume of the internally-pressurized chamber  180 . 
     During a rebound (outward) stroke the effects described above are reversed. In this case, since the piston rod  120  is being withdrawn, it occupies less of the total available internal volume of the damper cylinder  150 . The space vacated by the withdrawn piston rod  120  is filled by the hydraulic fluid  170  which is urged downward by the pressure above the floating piston  160  to fill the vacated space. In so doing, the floating piston  160  moves downward, increasing the volume of the internally-pressurized chamber  180  above it, which correspondingly reduces the pressure somewhat. 
     The above-described principles of operation for a conventional DeCarbon-type single-tube, pressurized damper such as shown in  FIG. 1  are well-known to those skilled in the art. 
     Referring now to  FIGS. 2 and 3 , additional structure in accordance with a first exemplary embodiment of the present invention is shown added to the prior-art damper  100  of  FIG. 1 . Since the structure and function of several of the parts in  FIGS. 2 and 3  are substantially identical to those in  FIG. 1 , the corresponding parts are designated by the same reference numbers as in  FIG. 1 . (This also generally applies to all other FIGS. which follow.) A partition  210  is secured within the bore of the damper by a partition retaining ring  211 . This partition  210  physically divides the hydraulic fluid into one portion above the partition  210 , and another portion below it, thereby forming a reserve chamber  212 . The partition  210  has a plurality of rebound flow ports  220  covered by a check valve  230  which is lightly biased in contact with the partition  210  by a relatively soft check valve spring  231 . Additionally, the partition  210  has a central compression flow port  240  which, in the position illustrated in  FIG. 3 , is blocked at its upper end by the small end of an intensifier piston  250 . 
     The intensifier piston  250  is located within an intensifier housing  260 , which can be integral with the damper cylinder  150  (as shown), or can be a separate structure sealed and retained within the bore of the damper cylinder  150 . During upward movement of the intensifier piston  250  as occurs during operation (to be described in detail further on), the intensifier piston  250  is prevented from exiting the intensifier housing  260  by the intensifier retaining ring  251 . The intensifier piston is sealingly engaged with the intensifier housing  260  at its upper (large diameter) end, as well as at its lower (smaller diameter) end. There is at least one vent port  270  which vents the space  214  between the upper and lower seals of the intensifier piston  250  to outside atmospheric pressure. There is also at least one bi-directional flow port  280  which passes vertically through intensifier housing  260 . 
     Still referring to  FIGS. 2 and 3 , the principles of operation of the present embodiment are described in the following paragraphs. 
     During a rebound stroke, the piston rod  120  is withdrawn from the damper cylinder  150 , resulting in some amount of vacated volume toward the lower end of the damper cylinder  150 . As described previously, this results in downward movement of the floating piston  160 , as well as a downward flow of the hydraulic fluid  170  immediately below it. Since downward movement of the floating piston  160  reduces the space between the floating piston  160  and the partition  210 , and since hydraulic fluid is incompressible, hydraulic fluid flows down through the bi-directional flow port(s)  280 . It then flows down through the partition  210  via the rebound flow port(s)  220 . It does this by opening the check valve  230  against the relatively light resistance of the check valve spring  231 . 
     During a compression stroke, the piston rod  120  and the damping piston  140  move further into the damper cylinder  150 , thus displacing a volume of the hydraulic fluid  170  equal to the volume of the additional length of the piston rod  120  which enters the damper cylinder  150 . As described previously, this results in an upward flow of the displaced volume of hydraulic fluid, accommodated by an upward movement of the floating piston  160 , which somewhat decreases the volume, and increases the pressure, in the internally-pressurized chamber  180 . However, in order to do so, the displaced volume of hydraulic fluid must first pass through the partition  210 . To achieve this, the fluid must create an upward force (pressure) at the lower (small) end of the intensifier piston  250  which is sufficient to overcome the downward force (pressure) at the upper (large) end of the intensifier piston  250 . To do so requires a pressure at the lower end of the intensifier piston  250  that is greater than the pressure at the upper end of the intensifier piston  250  by a multiple approximately equal to the ratio of the cross-sectional area of the large end of the intensifier piston  250  to the cross-sectional area of the compression flow port  240 . 
     For simplicity, it is assumed that the diameter of the small end of the intensifier piston  250  is only slightly greater than the diameter of the compression flow port  240 . Thus, the annular contact area between these parts is relatively quite small, and it can be said that, for flow through the compression flow port  240 , a pressure is required at the lower end of the intensifier piston  250  that is greater than the pressure at the upper end of the intensifier piston  250  by a multiple approximately equal to the ratio of the area of its large end divided by the area of its small end. 
     This pressure differential (multiple) between the small end and large end of the pressure intensifier  250  creates a compression damping effect in the damper. 
     Here is an example. Assume the diameter of the large end of the intensifier piston  250  is twice the diameter of the small end, and thus that the ratio of their cross-sectional areas is 4:1. Assume the diameter of the piston rod  120  is O½″, and thus it has a cross-sectional area of about 0.2 square inches. Assume the damping piston  140  has traveled inward into the damper cylinder  150  some distance (i.e., it is not fully-extended or “topped-out” against the seal head  130 ), as shown in  FIG. 2 . Assume that the pressure of the internally-pressurized chamber  180  above the floating piston is 100 psi. Assume static conditions, with the damping piston  140  not moving. Given these assumptions, and based on elementary principles, there is a uniform pressure of 100 psi throughout the interior of the damper. Furthermore, it can be readily calculated that, under these static conditions, the 100 psi internal pressure acting on the 0.2 square inch cross-sectional area of the piston rod  120  creates a 20-pound force tending to extend the piston rod  120 . In racing circles, this 20-pound force is sometimes referred to as “static nose force”. 
     The above described static conditions. Now the compression damping effect produced by the intensifier piston  250  during a compression stroke (inward movement of the piston rod  120 ) is described. Per basic principles, for an intensifier piston  250  with a cross-sectional area ratio of 4:1, a pressure of approximately 400 psi at the small end is required to overcome the 100 psi pressure at the large end (which originates from the internally-pressurized chamber  180  above the floating piston  160 ), in order to cause the intensifier piston  250  to move upward, thus unblocking the compression flow port  240  and allowing upward flow of the hydraulic fluid  170  displaced by the inward movement of the piston rod  120 . 
     For simplicity, it is assumed in the following discussion that the damping piston  140  has several large thru-holes and no restrictive valving (note that, actually, the exemplary embodiments of the present invention generally do incorporate restrictive valving on the damping piston  140  which does create compression damping forces). In other words, for purposes of clarity in describing the basic principles of the present embodiment, it is assumed here that the damping piston  140  itself creates no compression damping forces. Now, the 400 psi pressure created at the small end of the intensifier piston  250  acts uniformly throughout all portions of damper cylinder  150  below the intensifier piston  250 . Acting on the 0.2 square inch cross-sectional area of the piston rod  120 , it creates an 80-pound “dynamic nose force”. The difference between the previous 20-pound “static nose force” and this 80-pound “dynamic nose force” is 60 pounds; this 60 pounds represents the compression damping force produced by the present embodiment. Increasing the diameter and cross-sectional area of the piston rod  120 , of course, would create an even greater damping force. 
     To further describe the principles of the present embodiment, in the following it will be assumed that the above compression stroke continues inward for a distance sufficient to move the floating piston  160  upward some amount and increase the pressure in the internally-pressurized chamber  180  from 100 psi to 150 psi. This 150 psi pressure, of course, acts on the large end of the intensifier piston  250  and now approximately 600 psi pressure (basic 4:1 ratio) is required at the small end of the intensifier piston  250  in order for it to remain open, allowing continuation of the compression stroke. With 600 psi now acting on the 0.2 square inch cross-sectional area of the piston rod  120  a 120-pound “dynamic nose force” is now produced. In other words, as the compression stroke continues and the damping piston  140  and piston rod  120  travel further into the damper cylinder  150 , the volume of hydraulic fluid displaced by the piston rod  120  causes the floating piston  160  to move upward, which increases the pressure in the internally-pressurized chamber  180 , which increases the compression damping effect produced by the present embodiment. 
     Put another way, the present embodiment produces a “position-sensitive” compression damping effect, with the compression damping force increasing as the piston rod  120  and the damping piston  140  move further into the damper cylinder  150 . The extent and degree of this position-sensitive effect is influenced by the pre-set volume of the internally-pressurized chamber  180  above the floating piston  160 , relative to the diameter and maximum available travel of the piston rod  120 . If the pre-set volume of the internally-pressurized chamber  180  is relatively large, the position-sensitive effect is reduced. If the pre-set volume is relatively small, the position-sensitive effect is increased. 
       FIGS. 4 ,  5 , and  6 , show another exemplary embodiment of the present invention. This embodiment differs from the previous embodiment of  FIGS. 2 and 3  primarily due to an alternate configuration of the intensifier piston  255 , as best seen in  FIG. 5 . As compared with the previous “solid” intensifier piston  250  of  FIG. 3 , the intensifier piston  255  of  FIG. 5  has an intensifier piston compression flow port  256  which passes through its center. Another difference is the addition of an intensifier bleed screw  257  instead of the vent port  270  in  FIG. 3 . During assembly of the intensifier piston  255  into the partition  262 , this feature enables any trapped air pressure in the space between the upper and lower seals of the intensifier piston  255  to be vented by removing the intensifier bleed screw  257 . It further enables said space to be sealed off again, to provide proper operation, by re-installing said screw. This is done as part of the final assembly of these components. 
     Still referring to  FIG. 5 , the intensifier retaining ring  258  utilized here differs in form, but not function, from the previous intensifier retaining ring  251  of  FIG. 3 , Similarly, the check valve  235 , the check valve spring  236 , and the rebound flow port  222  as shown in  FIG. 5  all differ in form, but not function, from the equivalent features illustrated in  FIG. 3 . 
     One practical advantage of the embodiment of  FIG. 5  as compared with the embodiment of  FIG. 3  is that it combines the functions of both the partition  210  and the intensifier housing  260  of  FIG. 3  into one component, the partition  262  of  FIG. 5 . This reduces total part count and cost of the damper unit. 
     In operation during a compression stroke, fluid displaced by inward movement of the piston rod  120  applies pressure to the small end of the intensifier piston  255  via the arc flow port(s)  245 . Similar to the principles of operation of the previous embodiment, the intensifier piston  255  moves upward to permit upward flow of hydraulic fluid when the pressure ratio between the small end and the large end equals the area ratio of the large and small ends. For the intensifier piston  255  of  FIG. 5 , this statement refers specifically to the annular areas of the large and small ends. 
     Here is an example. Assume the ratio of the annular area at the large end of the intensifier piston  255  to the annular area at the small end is 2:1. Also assume that the nitrogen in the internally-pressurized chamber  180  above the floating piston  160  exerts a downward pressure of 100 psi on the annular area at the large end of the intensifier piston  255 . Given these parameters, and in accordance with basic principles, a pressure of 200 psi must be applied to the annular area at the small end of the intensifier piston  255  in order to cause the intensifier piston  255  to move upward and permit upward flow of the displaced hydraulic fluid through the intensifier housing arc port(s)  245 , and then up through the intensifier piston compression flow port  256 . 
       FIG. 7  illustrates the shaft displacement rebound fluid flow  270  that occurs through the structure of  FIG. 5  during a rebound stroke of the damper. Similarly,  FIG. 8  illustrates the shaft displacement compression fluid flow  271  that occurs during a compression stroke of the damper. 
       FIG. 9  shows another exemplary embodiment of the present invention. This embodiment is similar to the previous embodiment shown in  FIGS. 4 ,  5 ,  6 ,  7 , and  8 , except that a remote reservoir assembly  310  has been added. Also, the intensifier assembly  330  has been moved upward to the upper end of the damper cylinder  150 . The remote reservoir assembly  310  is connected to the main damper cylinder assembly  320  by an hydraulic hose  340 . The remote reservoir assembly  310  includes a reservoir end fitting  312 , a reservoir cylinder  314 , a floating piston  160 , and a reservoir cap  316 . 
     One advantage of the embodiment of  FIG. 9  as compared with previous embodiments is that, for a given length of damper cylinder  150  it increases the available travel distance of the damping piston  140  (available “damper stroke”). 
       FIGS. 10 and 11  show a exemplary embodiment of the present invention comprising a piggyback eyelet  410  with an attached reservoir cylinder  314  containing a floating piston  160  and an intensifier assembly  420 . The function of the partition  421 , the check valve  422 , the check valve spring  423 , and the rebound flow port  424  of this embodiment are similar to the corresponding structures of the previous embodiment shown in  FIG. 3 . The compression flow port  425  in the partition  421  provides a compression flow path for fluid from the damper cylinder  150  to an upward-facing annular area of the intensifier piston  426 . Due to the piston seal  427  and the vent  428  provided, the two other upward-facing areas of the intensifier piston  426  are at atmospheric pressure (considered zero pressure for purposes of this description). The large area of the bottom face of the intensifier piston  426  is subjected to the pressure within the internally-pressurized chamber  180  below the floating piston  160 . The intensifier piston  426  is fitted with an intensifier retaining ring  429  to ensure that it remains within the partition  421  during assembly and other possible conditions. 
     Similar to the principles of operation described for previous embodiments of the present invention, under static conditions the intensifier piston  426  is urged upward by the pressure on its bottom face into firm, sealing contact with the partition  421 . The intensifier piston  426  remains in firm sealing contact with the partition  421  unless the fluid pressure from the compression flow port  425  exerted downward against the upward-facing annular area  430  of the intensifier piston  426  creates sufficient force to overcome the upward force exerted by pressure on the bottom face of the intensifier piston  426 . This requires that pressure in the compression flow port  425  equals a multiple of the pressure in the internally-pressurized chamber  180 ; said multiple being approximately equal to the ratio of the area of the bottom face of the intensifier piston  426  to the area of the upward-facing annular area  430  of the intensifier piston  426 . 
     The relationship noted above is approximate only, due to the relatively narrow annular overlap area where the intensifier piston  426  contacts the partition  421 . During operation, when the intensifier piston  426  moves downward, and downward compression fluid flow occurs, the compression fluid pressure acting downwardly on this generally narrow annular edge portion of the overall upward-facing annular area  430  of the intensifier piston  426  is somewhat reduced in accordance with Bernoulli principles. 
     Similar also to previous embodiments: the increased pressure that is required to urge the intensifier piston  426  downward, to permit flow of the displaced fluid, acts on the cross-sectional area of the piston rod  120 , thus creating a compression damping force. 
       FIGS. 12 and 13  show a modified version of the embodiment of  FIGS. 10 and 11  which provides external adjustability of the compression damping force produced by the intensifier assembly  440 . This modified embodiment includes an intensifier adjusting screw  441 , an adjuster piston  442 , and an adjuster coil spring  443 . In operation, rotation of the intensifier adjusting screw  441  increases or decreases the preload force on the adjuster coil spring  443 . This force is transmitted through the adjuster piston  442  as an increased or decreased pressure in the adjacent hydraulic fluid  445 . This increased or decreased pressure is communicated to the upward-facing areas of the intensifier piston  446  with which the hydraulic fluid  445  is in contact. The downward force thus created on the intensifier piston  446  reduces, to a greater or lesser degree depending on the specific adjustment of the preload force on the adjuster coil spring  443 , the compression fluid pressure required to cause the intensifier piston  446  to move downward to permit compression fluid flow. Thus, this adjustment mechanism alters the compression damping force which is experienced at the piston rod  120 . 
       FIGS. 14 and 15  show another exemplary embodiment of the present invention. This embodiment utilizes an intensifier assembly  460  structure similar to that of  FIGS. 2 and 3 , but incorporated into the upper end of a reservoir cylinder  314  similar to that of  FIGS. 10 and 11 . The principles of operation for this embodiment are identical to those previously described for  FIGS. 2 and 3 . 
       FIGS. 16 and 17  show yet another exemplary embodiment of the present invention. This embodiment utilizes an intensifier assembly  510  similar to that of  FIGS. 14 and 15 , but, in addition, provides external adjustability via an intensifier adjuster knob  512 . The principles of operation for this embodiment are similar to those previously described for  FIGS. 14 and 15 , except for operation of the adjuster structure which is described in the following. 
     As best seen in  FIG. 17 , an external rotatable intensifier adjuster knob  512  is secured to a freely-rotating hex driver shaft  514  which includes a downwardly-projecting male hex portion which is keyed into a female hex portion of a threaded spring base  516  which rotates with it. The intensifier adjuster knob  512  is fitted with at least one detent ball  518  and one detent spring  519  which provide a detent function by providing audible and tactile feedback for each quarter turn (for example) adjustment of the intensifier adjuster knob  512 , as well as by helping to secure it at any pre-set position. The threaded spring base  516  is threaded on its outside diameter to produce axial movement upon rotation. Depending on the direction of rotation of the intensifier adjuster knob  512 , axial movement of the threaded spring base  516  increases or decreases the spring preload force of the intensifier adjuster spring  520 . 
     The principles of operation of this adjustment are described in the following. 
     First, as previously described, the basic principle of operation of the intensifier piston  522  itself can be best characterized as: in order for the intensifier piston  522  to move downward (“open”), the force(s) acting downward on the small end of the intensifier piston  522  must equal (or, actually, slightly exceed) the force(s) acting upward on the big end. For the embodiment as shown in  FIGS. 16 and 17 , the force acting upward on the big end of the intensifier piston  522  equals the cross-sectional area of the big end times the pressure in the internally-pressurized chamber  180 . Next, as to the small end of the intensifier piston  522 , there are two forces acting downward on it. One force is the compression fluid flow pressure acting on the small end of the intensifier piston  522  times the cross-sectional area of the small end. The other force is the force exerted by the intensifier adjuster spring  520 . These two forces together must slightly exceed the upward force on the big end of the intensifier piston  522  for the intensifier piston  522  to move downward (“open”), permitting compression fluid flow. 
     In accordance with the above principles, turning the intensifier adjuster knob  512  to increase the preload force of the intensifier adjuster spring  520  reduces the compression damping force effect produced by the adjustable intensifier assembly  510 . In fact, depending on specific parameters including the spring constant (“stiffness”) of the intensifier adjuster spring  520 , it would be possible to adjust for enough spring preload force to pre-set the intensifier piston  522  in an initially “open” condition such that the adjustable intensifier assembly  510  produced no flow restriction, and thus no compression damping force. Extending this example, a combination of parameters could be determined according to this embodiment of the present invention such that the pressure build-up in the internally-pressurized chamber  180  at some pre-determined point in the compression travel (“stroke”) of the piston rod  120  exceeded the spring preload force, thus closing the intensifier piston  522  and thus creating a compression fluid flow restriction and a compression damping effect. In other words, a combination of parameters could be chosen whereby the compression damping force produced varied from zero for the first portion of a compression stroke, to a finite and increasing value beyond that first portion. 
     Conversely, turning the intensifier adjuster knob  512  to decrease the preload force of the intensifier adjuster spring  520  increases the compression damping force effect produced by the adjustable intensifier assembly  510 . 
       FIGS. 18 and 19  show another exemplary embodiment of the present invention. This embodiment incorporates an intensifier piston  540  and partition assembly  550  similar in structure and function to that previously described in  FIGS. 2 and 3 . However, the key difference here is that, in  FIGS. 18 and 19  the pressure acting on the large end of the intensifier piston  540  is supplied by an external pressure source (not shown), not by an internal pressure source, such as the internally-pressurized chamber  180  as it was in previous embodiments. Thus, the pressure required at the small end of the intensifier piston  540  to permit compression fluid flow, and therefore the compression damping force produced, depends on the external pressure supplied. The pressure in  FIGS. 18 and 19  is supplied to the externally pressurized chamber  560  through a pressure port  562  fed by an external source (not shown) via a pressure fitting  564 . The pressure source, and the medium contained in the externally pressurized chamber  560  can be either pneumatic or hydraulic. A pneumatic medium and system is preferred where simplicity and low cost are dominant factors. An hydraulic medium is preferred where rapid responsiveness (quick response times) is important. 
     As shown in  FIG. 19 , a pressure chamber sealing head  566  is held in place by seal head retaining rings  568 , and seals the upper end of the externally pressurized chamber  560 . 
     One advantage of the embodiment of  FIGS. 18 and 19  is the remote, external controllability provided. A system could be designed, for example, utilizing various sensors on a vehicle. The information from these sensors, could be input to an on-board computer module having a pre-established algorithm for determining, for any given combination of inputs, the amount of pressure to be applied to the externally-pressurized chamber  560 , and, thus, the desired level of compression damping produced by the damper. A system of this type, utilizing an hydraulic medium, could sense actual vehicle conditions and respond within milliseconds of real-time, providing enhanced dynamic performance. 
       FIGS. 20A and 21  show another exemplary embodiment of the present invention. This embodiment is similar to the embodiment of  FIGS. 18 and 19  except that the externally pressurized chamber  560  is directly pressurized by the spring force of a suspension spring, such as a coil-over spring  570 . The upper end of the coil-over spring  570  is supported by a first portion  572   a  of a moveable element, such as a spring support ring  572 . The lower end of the coil-over spring  570  (not shown) is supported by a ring (not shown) attached to the lower eyelet (not shown, but equivalent to lower eyelet  112  in  FIG. 1 ). The spring support ring  572  has a second portion  572   b  in sealed, slidable contact with the damper cylinder  150  and the support ring housing  574 . The space between the spring support ring  572  and the support ring housing  574 , as well as the space in the externally pressurized chamber  560 , is filled with hydraulic fluid. Note that this hydraulic fluid is entirely distinct and separated from the hydraulic fluid contained within the rest of the damper unit. 
     The principles of operation of the embodiment of  FIGS. 20A and 21  are similar to those described for the embodiment of  FIGS. 18 and 19 . The only difference is that in  FIGS. 20A and 21  the pressure source is the external coil-over spring  570 , rather than a generalized pressure source. In a typical implementation, the compression damping force produced by the intensifier assembly  580 , from the beginning to the end of a full-travel compression stroke, would begin at a level determined by the initial preload of the coil-over spring  570 , then increase linearly with the depth of the compression stroke, according to the spring rate (“stiffness”) of the coil-over spring  570 . This characteristic could be described as a linearly-increasing position-sensitive compression damping curve. 
     In other words, assuming a typical “linear” coil-over spring  570 , the compressed force of the coil-over spring  570  would increase linearly as it was compressed (i.e., decreased in length). This force, directly supported by the spring support ring  572 , would produce a pressure in the externally pressurized chamber  560  that varied in direct proportion. This pressure, multiplied by the intensifier piston  540 , would proportionally increase the required pressure to unseat the small end of the intensifier piston  540  to permit compression fluid flow, and thus would proportionally increase the compression damping force produced as a function of the depth of the compression stroke. 
       FIG. 20B  shows an alternate version of the embodiment of  FIG. 20A , including addition of a secondary spring  576  in series with the main coil-over spring  570 , a dual-spring adaptor ring  577 , and a travel limit retainer ring  579 . The location of the travel limit retainer ring  579 , and the spring rate of the secondary spring  576  relative to the main coil-over spring  570 , is determined such that, during a compression stroke of the damper, the spring adaptor ring  577  engages the travel limit retainer ring  579  at some selected point in the travel. For example, on a damper with a maximum available stroke of 4-inches, the spring adaptor ring  577  might engage the travel limit retainer ring  579  at mid-stroke (i.e., at 2-inches of travel). In this example, the spring force supported by the spring support ring  572 , would increase linearly for the first 2-inches of travel, as the secondary spring  576  continued to compress. However, due to the function of the travel limit retainer ring  579 , for travel beyond this point the secondary spring  576  does not compress any further (only the main coil-over spring  570  continues to compress), and thus the spring force supported by the spring support ring  572  does not increase beyond the first 2-inches of travel. 
     Still referring to  FIG. 20B , in accordance with principles previously described, the compression force transmitted from the first portion  572   a  of the spring support ring  572  to the second portion  572   b  of spring support ring  572  produces a pressure in the externally pressurized chamber  560  that varies in direct proportion with the compression force. In turn, this pressure, multiplied by the intensifier assembly  580 , proportionately increases the required pressure to unseat the small end of the intensifier piston  540  to permit compression fluid flow. Thus, the compression damping force produced by the intensifier assembly  580  as a function of the depth of the compression stroke has the following characteristic shape: it begins at a level determined by the initial spring preload (the force of both springs is equal until the travel limit retainer ring  579  is engaged), it then increases linearly with travel until the spring adaptor ring  577  engages the travel limit retainer ring  579 , at which point it remains constant (“flattens out”) regardless of increasing travel. This type of compression damping characteristic is desirable for certain applications. 
       FIG. 20C  shows another alternate version of the embodiment of  FIG. 20A , including addition of a secondary spring  578  in series with the main coil-over spring  570 , a dual-spring adaptor ring  577 , and a spring travel limit retainer ring  579 . In  FIG. 20C  the location of the travel limit retainer ring  579 , and the spring rate of the secondary spring  578  relative to the main coil-over spring  570 , is determined such that the adaptor ring  577  is initially in engagement with the travel limit retainer ring  579 . Generally, at initial conditions (full extension of the damper), the secondary spring  578  has significantly more preload force than the main coil-over spring  570 . Therefore, during the first portion of damper travel, only the main coil-over spring  570  compresses. 
     For example, on a damper with a maximum available stroke of 4-inches, the preload on the secondary spring  578  could be such that only the main coil-over spring  570  compresses for the first 2-inches of travel. The spring force supported by the spring support ring  572 , would remain constant for the first 2-inches of travel. However, in this example, beyond this point the secondary spring  576  would begin to compress further (both springs compress), and thus the force supported by the spring support ring  572  would increase beyond the first 2-inches of travel. In contrast to the compression damping characteristic described above for the embodiment of  FIG. 20B , the embodiment of  FIG. 20C  produces a characteristic shape as follows: it begins at a level determined by the initial preload of the secondary spring  578 . It remains constant at that level (“flat”) until the point is reached where the secondary spring  578  begins to compress further, at which point the compression damping force begins to increase linearly with travel. 
     By extending the general principles illustrated by  FIGS. 20A ,  20 B, and  20 C, other possible compression damping force vs. depth of compression stroke characteristics can be achieved. 
       FIGS. 22 and 23  show another exemplary embodiment of the present invention. This embodiment is similar to the embodiment of  FIGS. 20A and 21  except that the floating piston  160  (not included or shown in  FIGS. 22 and 23 ), as utilized in all previous embodiments, has been entirely eliminated. This is feasible with the embodiment of  FIGS. 22 and 23 , since the compressed force of the coil-over spring  570  acts as a pressure source on the fluid within reservoir chamber  591 , within the damper unit similar to that previously provided by the floating piston  160 . As in the exemplary embodiments of  FIGS. 20A ,  20 B, and  20 C, the compression force of spring  570  is transmitted from the first portion  572   a  of the spring support ring  572  to the second portion  572   b  of the spring support ring  572 , thereby producing a pressure in the reservoir chamber  560 . 
     In the embodiment of  FIGS. 22 and 23 , as compared with the embodiment of  FIGS. 20A and 21 , the same hydraulic fluid is utilized throughout the entire damper unit, including the intensifier assembly  590  portion. Also, since there is no floating piston and no compressible gas (nitrogen) in this embodiment, the fluid volume displaced by the piston rod  120  during a compression stroke must be accommodated by downward movement of the spring support ring  572 , thus providing additional annular volume for the displaced fluid. 
     It should be noted, that this also has the effect of somewhat increasing the “effective spring rate” of the coil-over spring  570 . For example, assume a coil-over spring  570  with a spring rate of 300 lbs/in. Also assume that the ratio of the annular area of the spring support ring  572  to the cross-sectional area of the piston rod  120  is 10-to-1. Assume further, for simplicity of this example, that the coil-over spring  570  is not at all compressed (has zero pre-load force) at full extension of the damper. Now assume a compression stroke that shortens the damper exactly 1-inch. Although the damper is only 1-inch shorter, the coil-over spring  570  is now 1.1-inches shorter. This results from the 1-inch damper stroke, plus the 0.1-inch downward movement of the spring support ring  572  to accommodate the fluid volume displaced by the piston rod  120 . Thus, the force exerted by the coil-over spring  570  in this position is 330 lbs, and it has an “effective spring rate” of 330 lbs/in. 
     One advantage of the embodiment of  FIGS. 22 and 23  is the complete elimination of the floating piston  160 , the internally-pressurized chamber  180 , and the Schrader valve  190 , as included in all previous embodiments. Another advantage, shared with the embodiment of  FIGS. 20A and 21  is the linearly-increasing, position-sensitive compression damping effect produced by the intensifier assembly  590 . 
     Note that, generally, the total compression damping force produced by the embodiment of  FIGS. 22 and 23 , as well as other embodiments of the present invention, will also include the non-linearly-increasing, non-position-sensitive compression damping forces produced by conventional compression valving at the damping piston  140 . Thus, the overall compression damping characteristics will be a combination of those produced at the damping piston  140 , plus those produced by the intensifier assembly  590 . 
       FIGS. 24 ,  25  and  26  show another exemplary embodiment of the present invention. This embodiment is similar to the embodiment of  FIGS. 22 and 23  except that an intensifier piston  610  similar to that first shown in  FIGS. 4 ,  5  and  6  is utilized. Another difference is the addition of the intensifier preload spring  612 . This enables an increase in the compression damping effect produced by the intensifier piston  610  near full extension, and less relative progressivity throughout the stoke, without requiring an increase in the spring preload of the main coil-over spring  570 . An optional small bleed orifice  614 , permitting limited fluid flow through the intensifier piston  610  when in a closed condition, thus modifying the operative characteristics of the intensifier assembly, is included. It should be noted that the bleed orifice  614  included here, although not illustrated other embodiments, could also be incorporated in them if desired. 
       FIG. 27  shows another exemplary embodiment of the present invention. This embodiment is similar to the embodiment of  FIGS. 24 ,  25  and  26  except that, rather than the previous intensifier preload spring  612  (as shown in  FIG. 25 ), an intensifier open-bias preload spring  618  is utilized. The effect of the intensifier open-bias preload spring  618  is to maintain the intensifier piston  616  in an open (no flow restriction) position during the early portion (i.e., near-full-extension portion) of a compression stroke. The intensifier piston  616  does not tend to close until a point in the compression stroke is reached where the internal pressure generated by the coil-over spring  570  overpowers the intensifier open-bias preload spring  618 . At this point, the intensifier assembly begins to produce a compression damping effect by requiring pressure at the small end of the intensifier piston  616  in order to keep it open. 
     A characteristic of having no compression damping created by the intensifier at near-full-extension, but with some beginning and increasing intensifier-created compression damping occurring somewhere mid-stroke can be desirable for certain applications. 
     Note that, by combining the general principles illustrated by  FIGS. 20A ,  20 B, and  20 C with those illustrated by  FIGS. 25 and 27 , a wide variety of possible compression damping force vs. depth of compression stroke characteristics can be achieved. 
       FIG. 28  shows an exemplary embodiment of the present invention as incorporated into the FLOAT-series of air-sprung dampers as produced by FOX Racing Shox of Watsonville, Calif. In this embodiment, an adjustable intensifier assembly  510  essentially identical to that previously shown in  FIG. 17  is attached to the main damper assembly  630  by the piggyback eyelet  632 . The pressurized air  640  for the air-sprung feature of the damper is supplied via the Schrader valve  642  as shown. 
     During a compression stroke of the damper, a volume of hydraulic fluid  170  displaced by the piston rod  620  flows upward via the central port  622  in the piston rod  620 , then flows to the right via a horizontal port  634  in the piggyback eyelet  632 , then flows downward via an angled port  636  into the intensifier assembly  510 . The horizontal port  634  is drilled or otherwise manufactured approximately on-axis with the Schrader valve  642 . A press-fit sealing ball  644  is pressed into the entrance of the horizontal port  634  in order to keep the hydraulic fluid  170  and the pressurized air  640  entirely separate. 
     One advantage of the embodiment of  FIG. 28  is that, by providing for flow of the displaced hydraulic fluid up through the piston rod  620  to reach the intensifier assembly  510  via ports in the piggyback eyelet  632  as shown, the pressure chamber sleeve  660  can be easily and conveniently unthreaded and completely removed downward from the overall assembly for the periodic cleaning and maintenance typically required to remove foreign matter which may pass through the dynamic seals during operation and accumulate over time. With a more conventional construction utilizing an attached reservoir at the bottom end of the damper assembly, removal of the pressure chamber sleeve  660  is significantly more difficult, since the pressure chamber sleeve  660  as shown cannot be removed in an upward direction due to interference between the chamber seal assembly  670  and the outer seal assembly  680  portion of the pressure chamber sleeve  660 . Thus, additional disassembly, or added complexity of construction, would be required to enable removal of the pressure chamber sleeve  660  if the reservoir was attached at the bottom end of the damper assembly. 
       FIG. 29  shows another exemplary embodiment of the present invention. Two of the unique features of this embodiment, as compared with all previously shown embodiments, are the outer sleeve  710 , and the seal head check valve assembly  720 . A third differentiating feature is the lack of compression valving (symbolic)  142  (not included or shown in  FIG. 29 ) as shown and identified in  FIG. 1 , and as illustrated in all previous embodiments. The partition  210  and the intensifier piston  730  are similar to those previously shown and described per  FIGS. 2 and 3 , except for the addition of a bleed screw  257  in the intensifier piston  730  for purposes as first previously described relative to  FIGS. 4 and 5 . This feature is important for the embodiment of  FIG. 29 , since a vent port  270  (not shown or included in  FIG. 29 ) feature such as shown in  FIGS. 2 and 3  would be difficult to achieve due to the added outer sleeve  710  of  FIG. 29 . 
     A primary advantage of the embodiment of  FIG. 29  is that, since the damping piston  140  has no compression ports or valves, no hydraulic fluid flows through the damping piston  140  during a compression stroke. Therefore, the displaced fluid volume during a compression stroke is determined by the full cross-sectional area of the damping piston  140 , rather than by the much smaller cross-sectional area of the piston rod  120 , as in previous embodiments. One portion of the displaced fluid, a portion equal to the displaced volume of the piston rod  120 , is accommodated by upward movement of the floating piston  160 . The other portion exits the damper cylinder  150  via the upper flow port(s)  741 , then travels downward via the annular space  742  between the damper cylinder  150  and the outer sleeve  710 , then re-enters the damper cylinder  150  via the lower flow port(s)  743 , which lead to the check valve assembly  720  in the seal head  130 . The check valve assembly  720  opens for flow in the upward direction, allowing the fluid flow to continue and to fill the vacated annular space behind the damping piston  140  during a compression stroke. Since there is no flow through the damping piston  140  during a compression stroke, the pressure generated by the intensifier piston  730  acts on the full cross-sectional area of the damping piston  140 . Thus, relatively large compression damping forces can be produced with this embodiment at significantly lower internal pressures than in previous embodiments. 
     In  FIG. 29 , the check valve assembly  720  permits fluid flow in the upward direction only. Thus, during a rebound stroke, the check valve assembly  720  is closed, and the fluid pressure created between the damping piston  140  and the seal head  130  cannot escape through the seal head  130 . Therefore, the desired rebound damping forces are created by the damping piston  140  and the rebound valving  141 . 
       FIGS. 30 and 31  show another exemplary embodiment of the present invention. This embodiment is somewhat similar to the previous embodiment shown in  FIGS. 16 and 17 , except that the intensifier assembly  750  is oriented horizontally within the piggyback eyelet  760  structure leading to the reservoir assembly  770 . Besides the different location, the other key difference relative to the embodiment of  FIGS. 16 and 17  is that here the intensifier adjuster spring  754  engages the large end of the intensifier piston  752 , rather than the small end. The net effect of this is that, in  FIGS. 30 and 31 , an adjustment that increases the preload force on the intensifier adjuster spring  754  increases the compression damping force produced by the intensifier assembly  750 . In contrast, in  FIGS. 16 and 17 , an adjustment that increases the preload force of the intensifier adjuster spring  520  decreases the compression damping force produced. There is no particular advantage or disadvantage to either construction; the differences are simply pointed out here for clarity. 
     In the embodiment of  FIGS. 30 and 31 , the pressure from the internally-pressurized chamber  180  below the floating piston  160  reaches the large end of the intensifier piston  752  via a pressure port  762  in the piggyback eyelet  760 . Fluid flow due to displacement of the piston rod  120  during compression and rebound strokes flows into and out of the upper portion of the reservoir cylinder  772  via the flow port  764 . 
       FIGS. 32 and 33  show another exemplary embodiment of the present invention. This embodiment differs from the previous embodiment of  FIGS. 30 and 31  in two basic respects. First, the intensifier assembly  780 , rather than being oriented horizontally as in the previous embodiment, is oriented at a small angle from horizontal. This provides no significant performance benefits, but is shown simply as an illustration of one of the configuration possibilities available with this embodiment which may offer easier access to the intensifier adjuster knob  512  for making adjustments to the damper as installed in a particular application. 
     Secondly, the embodiment of  FIGS. 32 and 33  differs from the previous embodiment of  FIGS. 30 and 31  by the addition of the compression flow bleed adjuster assembly  790 . The basic mechanism of this assembly, whereby rotation of the bleed adjuster knob  792  produces translation of the tapered bleed adjuster needle  794 , is similar to the mechanism utilized in the adjustable intensifier assembly  510 , as best seen and described previously relative to  FIG. 17 . The compression flow bleed adjuster assembly  790  provides independent tuning of compression bleed flow of the damper. This can be an important tuning element in many damper applications. Compression bleed flow occurs in parallel with any compression flow through the intensifier assembly  780 . 
       FIG. 34  shows an overall view of a front suspension fork  800  which could be used on a bicycle or motorcycle (not shown).  FIGS. 35 ,  36 , and  37  show an exemplary embodiment of the present invention as incorporated into the suspension fork  800  of  FIG. 34 . 
       FIGS. 35 and 36  show a fork leg assembly  810 , in part comprised of a fork crown (partial view)  812 , a fork upper tube  814 , a fork lower tube  816 , a Schrader valve  819 , air  820 , and hydraulic fluid  830  filled to an approximate level  831  as shown. In addition, the fork leg assembly  810  comprises a damper assembly  840  as shown in isolation in  FIG. 36 . The upper portion of the damper assembly  840  shown in  FIG. 36  includes a piston rod  842 , a damping piston  844 , a damper cylinder  850  and hydraulic fluid  830 , a construction sometimes referred to as a damper cartridge assembly. 
     An intensifier assembly  860  in the lower portion of the damper assembly  840  shown in  FIG. 36  comprises an exemplary embodiment of the present invention, and is best seen in  FIG. 37 . 
       FIG. 37  shows the intensifier assembly  860  which includes a partition  870 , an intensifier housing  880 , an intensifier piston  890 , an intensifier preload spring  892 , an adjuster rod  894 , and an adjuster knob  896 . 
     The principles of operation of the intensifier assembly  860  of  FIG. 37  are similar to those previously shown and described for previous embodiments. During a compression stroke of the suspension fork  800 , the piston rod  842  displaces a volume of the hydraulic fluid  830  in the damper cylinder  850 . In order for the compression stroke to occur, the displaced fluid must exit the damper assembly. For the structure shown in  FIG. 37 , this can only occur when the pressure in the hydraulic fluid  830  above the partition  870 , acting on the area of the small end of the intensifier piston  890 , overcomes the upward forces acting on the intensifier piston  890 , thus causing the intensifier piston  890  to move downward, allowing downward fluid flow through the compression flow port  872 . 
     There are two upward forces acting on the intensifier piston  890 . First, there is the upward force applied by the intensifier preload spring  892 . Second, there is the internal pressure in the air  820 , which is communicated by the hydraulic fluid exterior to the damper cylinder  850  up through the bottom of the intensifier assembly  860  via the lower bi-directional flow port(s)  862 , and which acts on the cross-sectional area of the large end of the intensifier piston  890  to produce the second upward force. 
     Thus, identical in principle to previously-described embodiments of the present invention, a relatively large hydraulic fluid pressure increase is created by the adjustable intensifier assembly  860  during a compression stroke. This pressure increase, acting on the cross-sectional area of the piston rod  842  produces a compression damping force in the suspension fork  800 . 
     The fork leg assembly  810  of  FIG. 35  can be assembled with a desired volume of air  820  at atmospheric pressure, or it can be supplied with pressurized air (or other compressible gas, such as nitrogen) via a Schrader valve  819 . In either case, as a compression stroke of the suspension fork  800  proceeds, the volume of the air  820  in the fork leg assembly  810  is progressively reduced (compressed), resulting in a progressively-increasing internal pressure. This increasing internal air pressure, acting through the intensifier assembly  860 , produces a progressive increase in the compression damping force of the suspension fork  800 . Thus, a progressive, position-sensitive compression damping force is produced. 
     However, it should be noted again that, similar to descriptions regarding previous embodiments of the present invention, compression damping forces in the suspension fork  800  are generally also produced at the damper piston  844 . Thus, in general, the total compression damping characteristics produced by various embodiments of the present invention result from a combination of the compression damping forces created by valving at the damper piston (for example,  844  in  FIG. 36 ) plus the compression damping forces resulting from pressure increases produced by the intensifier assembly (for example,  860  in  FIG. 36 ) acting on the cross-sectional area of the piston rod (for example,  842  in  FIG. 36 ). 
     Although the present invention has been explained in the context of several exemplary embodiments, minor modifications and rearrangements of the illustrated embodiments may be made without departing from the scope of the invention. For example, but without limitation, although the exemplary embodiments described intensifier pistons with bleed or vent provisions to eliminate pressure in the space between the small and large ends of the intensifier pistons, the principles taught may also be utilized in damper embodiments without these provisions. In addition, although the exemplary embodiments were described in the context of vehicular applications, the present damper may be modified for use in non-vehicular applications where dampers may be utilized. Furthermore, it is contemplated that various aspects and features of the invention described can be practiced separately, combined together, or substituted for one another, and that a variety of combination and subcombinations of the features and aspects can be made and still fall within the scope of the invention. Accordingly, the scope of the present invention is to be defined only by the appended claims.