Abstract:
A reduced weight and repairable piston accumulator. The accumulator includes a load bearing metallic cylinder with removable end caps secured thereto with slip flanges for allowing repairability and for achieving the required cycle life. The cylinder serves as the surface on which the piston slides and is designed such that it sustains the axial stress induced by pressurization of the accumulator. A composite over wrapping is designed such that it sustains the stress in the hoop (radial) direction. A stress transitioning bushing can be provided for transitioning hoop stresses between the overwrap and the slip flange. When combined with the cylinder, the fibers of the composite wrap will not be placed in shear and thus will not fatigue in the same manner as some prior art designs.

Description:
RELATED APPLICATIONS 
     This application claims the benefit of U.S. Provisional Application No. 60/986,400 filed Nov. 8, 2007, which is hereby incorporated herein by reference. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates generally to a lightweight composite high pressure piston accumulator. 
     BACKGROUND OF THE INVENTION 
     Demand for lightweight accumulators is increasing, especially for mobile applications (e.g., aircraft, motor vehicles, etc.) where extra weight can reduce fuel efficiency. One example of a mobile application of an accumulator is in a hybrid powertrain for a vehicle. The term “Hybrid” generally refers to the combination of one or more conventional internal combustion engines with a secondary power system. The secondary power system typically serves the functions of receiving and storing excess energy produced by the engine and energy recovered from braking events, and redelivering this energy to supplement the engine when necessary. The secondary power system acts together with the engine to ensure that enough power is available to meet power demands, and any excess power is stored for later use. This allows the engine to operate more efficiently by running intermittently, and/or running within its most efficient power band more often. 
     Several forms of secondary power systems are known. Interest in hydraulic power systems as secondary systems continues to increase. Such systems typically include one or more hydraulic accumulators for energy storage and one or more hydraulic pumps, motors, or pump/motors for power transmission. Hydraulic accumulators operate on the principle that energy may be stored by compressing a gas. An accumulator&#39;s pressure vessel contains a captive charge of inert gas, typically nitrogen, which becomes compressed as a hydraulic pump pumps liquid into the vessel, or during regenerative braking processes. The compressed fluid, when released, may be used to drive a hydraulic motor to propel a vehicle, for example. Typically operating pressures for such systems may be between 3,000 psi to greater than 7,000 psi, for example. 
     As will be appreciated, since the accumulator stores energy developed by the engine or via regenerative braking processes, it plays an important role in achieving system efficiency. One type of accumulator that may be used is commonly referred to as a standard piston accumulator. In a standard piston accumulator, the hydraulic fluid is separated from the compressed gas by means of a piston which seals against the inner walls of a cylindrical pressure vessel and is free to move longitudinally as fluid enters and leaves and the gas compresses and expands. 
     The piston is typically made of a gas impermeable material, such as steel, that prevents the gas from mixing with the working fluid. Keeping the gas from mixing with the working fluid is desirable, especially in high pressure applications such as hydraulic hybrid systems, to maintain system efficiency and avoid issues related with removing the gas from the working fluid. 
     In order to maintain a sufficient seal, the dimensional tolerance at the interface between the piston and the inner wall of the cylinder is generally very close. Further, the pressure vessel typically must be extremely rigid and resistant to expansion near its center when pressurized, which would otherwise defeat the seal by widening the distance between the piston and cylinder wall. This has generally eliminated the consideration of composite materials for high pressure piston accumulator vessels like those used in a hybrid system, for example, as composite materials tend to expand significantly under pressure (e.g., about 1/10 of an inch diametrically for a 12 inch diameter vessel at 5,000 psi pressure). Furthermore, the need to assemble the cylinder with a piston inside traditionally requires that the cylinder have at least one removable end cap for use in assembly and repair, rather than the integral rounded ends that are more structurally desirable in efficiently meeting pressure containment demands with composite materials. Composite pressure vessels are not easily constructed with removable end caps. 
     As a result of the foregoing, standard piston accumulator vessels tend to be made of thick, high strength steel and are very heavy. Standard piston accumulators have a relatively high weight to energy storage ratio as compared to other types of accumulators (e.g., bladder-type accumulators), which makes them undesirable for mobile vehicular applications (as such increased weight would, for example, reduce fuel economy for the vehicle). Therefore, despite their potentially superior gas impermeability, conventional piston accumulators are largely impractical for vehicular applications. 
     Another known composite accumulator uses an aluminum liner for both the piston travel surface and main liner of the pressure vessel. This design eliminates the need to pressure balance a secondary liner (e.g. by pressurizing the space between the main and secondary liner), but suffers from low fatigue endurance. The low fatigue endurance is usually caused by the difficulty of getting the aluminum liner (or other thin metal liner) to properly load share with the composite. Without the addition of an autofrettage process, this type of accumulator will have exceptionally low fatigue life. With an autofrettage process, the liner will grow erratically along its length making an adequate piston seal on the trapped piston nearly impossible resulting in gas mixing with the working fluid. 
     As noted, a consideration for accumulators in hydraulic hybrid systems is repairability. As noted, composite bladder accumulators are difficult to construct with removable end caps that would allow repair/replacement of the bladder and/or seals. Thus, in the event of seal failure, the entire accumulator is inoperable and must be discarded. To the degree that lightweight composite accumulators have had low cycle requirements or have been used on equipment that replacement was acceptable (aircraft, military vehicles, etc.), the use of such non-repairable bladder accumulators has been an acceptable practice. Placing lightweight accumulators in systems that are more commercial in nature and in larger numbers, however, makes non-repairable accumulators both financially and environmentally unsound. 
     U.S. Pat. No. 4,714,094 describes a repairable piston accumulator in which the all of the stresses (e.g., axial and hoop) are designed to be sustained by a composite overwrap. As a consequence of making a large enough opening for repairability and maintaining a thin non-load bearing liner (or minimally load bearing liner), the required primary wrap angle of the composite becomes 55 degrees placing some shear stress into the composite fibers. The shear stress is an undesirable condition and requires a second circumferential wrap to compensate for the stress. Thus, while the accumulator is repairable, the design likely fails to give the fatigue characteristics demanded by current and future uses of lightweight hydraulic accumulators. 
     SUMMARY OF THE INVENTION 
     The present invention provides a reduced weight and repairable piston accumulator. The accumulator includes a load bearing metallic cylinder with removable end caps secured thereto with slip flanges for allowing repairability and for achieving the required cycle life. The cylinder serves as the surface on which the piston slides and is designed such that it sustains the axial stress induced by pressurization of the accumulator. A composite wrapping is designed such that it sustains the stress in the hoop (radial) direction. The wind angle of the composite wrap can be, for example, between about 75 and about 90 degrees. When combined with the cylinder, the fibers of the composite wrap will not be placed in shear and thus will not fatigue in the same manner as some prior art designs. 
     In an embodiment, the cylinder of the accumulator is open at one end. In an alternative embodiment, the cylinder may be open at both ends. An autofrettage process may be done and the cylinder bore finished machined after the autofrettage. This allows for close tolerance piston seal and longer fatigue life on the cylinder. A bushing transitions stresses from the relatively low modulus central portion of the cylinder to the relatively high modulus slip flange area. The bushing produces a significant improvement in fatigue life over threaded caps (e.g., caps threaded onto the cylinder ends) and also helps to achieve the required fatigue life for high pressure applications such as hybrid transmission systems. 
     Accordingly, an accumulator comprises a liner having an open end and a radially outwardly extending shoulder at the open end, a composite overwrap wrapped around the liner for carrying hoop stress applied to the liner, a cap for closing the open end of the liner, and a slip flange for connection to the cap with the shoulder of the liner trapped between the cap and the slip flange. 
     A stress transition bushing can be provided in an area of transition between the overwrap and the slip flange for transitioning hoop stress from the overwrap region to the slip flange. The bushing can be tapered, for example, such as along its axial length such that it has a greater radial dimension at an end nearest the shoulder of the liner the slip flange. The slip flange can include a counterbore, and the bushing can be received at least partially within the counterbore. The counterbore can be tapered along its axial length so as to have a greater radius at an end nearest the overwrap, for example. The bushing can also be at least partially overwrapped with the composite overwrap. An inner diameter of the slip flange can engage an outer diameter of the liner, and at least a portion of the inner diameter of the slip flange that engages the outer diameter of the liner can be tapered along its axial length. The liner can have a thickness of approximately 0.375 inches, for example, but virtually any thickness can be used with sufficient overwrapping. The bushing can be a steel or carbon composite bushing. The accumulator can further include a pressure balanced liner and/or a piston supported for sliding axial movement within the accumulator and forming separate chambers within the accumulator. 
     In accordance with another aspect, a method of making an accumulator comprises forming a liner with an open end and with a radially outwardly extending shoulder at the open end thereof, positioning a slip flange over the liner axially inwardly of the radially outwardly extending shoulder, and closing the open end by securing a cap to the slip flange such that the shoulder of the cylindrical liner is trapped between the slip flange the cap. The forming the liner can include machining the liner from a tubular blank such as a conventional steel bladder accumulator liner, for example. 
     Further features of the invention will become apparent from the following detailed description when considered in conjunction with the drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a cut-away perspective view of an exemplary accumulator in accordance with the invention. 
         FIG. 2  is a longitudinal cross-sectional view of the accumulator of  FIG. 1 . 
         FIG. 3  is an enlarged portion of  FIG. 2  showing an exemplary slip flange connection. 
         FIG. 4  is a cross-sectional view of an exemplary close-fit slip flange connection. 
         FIG. 5  is a cross-sectional view of an exemplary slip flange connection including a bushing. 
         FIGS. 6-13  are cross-sectional views of various other exemplary slip flange connections. 
     
    
    
     DETAILED DESCRIPTION 
     Turning now to the drawings in detail, and initially to  FIGS. 1 and 2 , an exemplary lightweight, high pressure and repairable accumulator  10  is illustrated. The accumulator  10  is generally an elongate structure having an opening  14  at one end for receiving a fitting for connection to a gas source, such as high pressure nitrogen, and an opening  16  at the opposite end for receiving a fitting for connection to a hydraulic fluid source, such as a pump of a hybrid transmission system. A piston  18  is supported within the accumulator  10  and is displaced axially during pressurization/depressurization of the accumulator  10 . 
     The accumulator  10  is made from fiber overwrap  22 , typically composed of carbon and glass fibers, for example, that is wrapped around a tubular load bearing high strength steel liner  24  that is preferably cylindrical and also commonly referred to as a cylinder or shell. As will be appreciated, a composite material generally consists of two or more phases on a macroscopic scale whose mechanical performance and properties are designed to be superior to those of the constituent materials acting independently. One phase is usually discontinuous, stiffer and stronger and is called reinforcement, whereas the weaker phase is continuous and is called the matrix. Various types of fiber reinforcement include Glass, Carbon, Aramid and Boron, for example. Typical matrix materials include Polymers (e.g., Epoxy, Polyester, Thermoplastics), Metals (e.g., Aluminum, magnesium) and Ceramics. 
     In general, the steel liner  24  is designed to sustain the axial stress developed under pressurization of the accumulator  10 , while the composite overwrap  22  is designed to sustain the radial stress, also sometimes referred to as hoop stress, developed during pressurization. The ratio of carbon and glass in the composite overwrap will vary with the wrap layer and/or particular design of the accumulator  10 . 
     The composite overwrap  22  is typically wrapped in a largely circumferential manner with a wind angle of, for example, between about 75 to about 90 degrees with respect to the longitudinal axis of the accumulator  10 , to provide a maximum of strength in the hoop stress direction and a minimal amount in the axial direction. The composite overwrap  22  in the illustrated embodiments is also thicker at the ends to reduce and/or prevent flaring of the ends of the steel liner  24 . In the illustrated embodiment, one end of the steel liner  24  is formed as a dome, while the opposite end is closed by a releasably securable domed cap  28 . Between formed dome end and the domed cap end  28  is the midsection M generally defined as the region of the liner  24  that is overwrapped. A pressure balanced liner  30 , which may be steel or aluminum and may have a thickness between about 0.125-0.250 inches for example, can also be optionally provided as shown. The piston  18  includes a seal (not shown) for sealing against the pressure balanced liner  30  or the steel liner  24  in the absence of a pressure balanced liner  30 . For example, a bidirectional seal can be used that can compensate for changes in diameter of the steel liner  24  that may occur under pressure. 
     With reference to  FIG. 3 , details of the connection between the domed cap  28  and steel liner  24  are illustrated. The steel liner  24  has a shoulder  32  machined or otherwise formed at an end thereof and adapted to be engaged by a slip flange  36  telescoped over the liner  24 . The slip flange  36  is preferably a unitary annular piece that can be telescoped over the liner  24  as shown, but may alternatively be multiple pieces connected together and/or separately to the domed cap  28 . The domed cap  28  has an integral boss  38  with a groove for receiving the shoulder  32  of the liner  24 , and for mating with a corresponding surface of the slip flange  36  such that the shoulder  32  is trapped between the cap  28  and the slip flange  36 . A seal  40  is also provided for sealing the sleeve  24  to the domed cap  28 . The domed cap  28  and slip flange  36  are secured together with suitable fasteners  42 , such as screws or bolts, for example. As will be described in more detail below, the slip flange connection provides a robust connection that not only permits removal of the domed cap  28 , but also is designed to gradually transition hoop stresses from the central portion M of the accumulator  10  to the slip flange  36  to avoid damaging the steel liner  24 . 
     The slip flange connection in  FIGS. 1-3  includes a stress transition bushing  46  received in a counterbore  50  of the slip flange that is interposed between the slip flange  36  and the steel liner  24  for gradually transitioning stresses through the slip flange  36 . The bushing  46  can be a steel or carbon composite bushing, for example, and may be tapered and/or shaped so as to provide a gradual transition for the less stiff region to the right of the slip flange  36  in  FIG. 3 , to the more stiff region of the slip flange  36 . Similarly, the counterbore  50  can be shaped to achieve a similar effect, as will be described. 
     Turning now to  FIGS. 4-13 , various exemplary embodiments of the slip flange connection will be described. Each of the following exemplary embodiments tends to reduce the concentration of bending stresses in the steel liner  24  that may occur due to bending moments generated during pressurization of the accumulator adjacent the slip flange  36 . The concentration in bending may be exacerbated by sealing the bore at the right end, eliminating any pressure load outboard of the seal. 
       FIG. 4  illustrates a simple close slip fit or minor interference fit slip flange connection. In this embodiment, the slip flange  36  engages, along its axial length, the outer diameter surface of the liner  24 . No bushing is used, and high tensile fatigue stresses may occur on the inside in some applications if the slip flange bore is not tapered. To reduce such fatigue stresses, the slip flange bore can be tapered such that its diameter is greater on the side closer to the overwrap  22 , thereby allowing more expansion approaching the left face of the slip flange  36 . Such taper is represented in  FIG. 4  by dotted line T. For simplicity, the domed cap  28  is only being shown in  FIG. 4 . 
       FIG. 5  illustrates a tapered steel bushing  52  adjacent the slip flange  36  and under the overwrap  22 . The steel bushing  52  is generally retained by the overwrap  22  and gradually transitions stresses between the relative stiff slip flange  36  to the more compliant composite midsection region M. The bushing  52  is subject to high hoop fatigue stresses at its thin edge, so it typically will be made from high-strength-steel and finished well. The gap between the slip flange and the bushing and composite overwrap  22  may breathe during cycling. 
       FIG. 6  illustrates a slip flange  36  having a tapered transition section  56  formed therewith as an integral piece. This embodiment is similar to the embodiment of  FIG. 5  except that the bushing  52  of  FIG. 5  is essentially part of the slip flange  36  of  FIG. 6 . This design would generally eliminate any tendency for the joint to breathe. 
       FIG. 7  illustrates a carbon composite wrapped bushing  60  in a tapered counterbore  50 . Alternatively, the bushing  60  may be tapered in a straight counterbore  50 . In either case, the bushing  60  is interposed between the overwrap  22  and the steel liner  24 . By varying the clearance along the axial length of the bushing  60 , the stiffness can be transitioned from high at the right end to lower at the left. This embodiment may require precision machining. In order to sustain the potentially very high compressional loads, the wrapped bushing  60  can be made from a bidirectional composite in order to resist cracking under the compressive loads. 
       FIG. 8  illustrates a slip flange  36  having a slanted face  64  for engaging a corresponding angled face  66  on the shoulder  32  of the steel liner  24 . The forward slant face  64  tends to rotate the liner shoulder  32  to the right and reduce the stress concentration at the slip flange  36  and steel sleeve  24 . A bushing  68 , such as any one of the herein described bushing, can be used as shown. Alternatively, a close-fit design such as the design of  FIG. 4  can be used. 
       FIG. 9  illustrates a tapered carbon bushing  72  in a loose-fit counterbore of slip flange  36 . The bushing  72  provides a transition in stiffness without the close machining of the design of  FIG. 7 . 
       FIG. 10  illustrates another slip flange connection wherein the slip flange bolts holes are angled to bring their centerline closer to the applied pressure loads. This design typically will reduce the moments in the slip flange  36  by moving the stress concentration point from the flange corner to the flange edge, but manufacturing would be considerably more complicated. Any of the bushing designs disclosed herein could also be used in connection with this embodiment as well. 
       FIG. 11  illustrates a long tapered steel bushing  76  partially received in the slip flange counterbore  50 . The bushing  76  extends axially from the counterbore of the slip flange  36 , and increases the length of transition without adding weight (for example, compare to the bushing of  FIG. 5 ). 
       FIG. 12  illustrates a slip flange  36  having a slanted face  80  for engaging a corresponding angled face  82  on the shoulder  32  of the steel liner  24  in a dovetail fashion. 
       FIG. 13  illustrates a combination of the designs of  FIGS. 8 and 10 . 
     In the forgoing designs including a bushing, the bushing can have any suitable taper angle such as, for example, between about 15 and about 25 degrees. 
     It will be appreciated the accumulator  10  of the present invention is not only significantly lighter than equivalent sized steel designs, it is also repairable. The reduction in weight is generally made possible by relying on a thinner steel liner  24  combined with composite overwrapping, while the slip flange connection between the steel liner  24  and the domed cap  28  provides a robust yet releasably securable manner connecting the two parts. An accumulator of the present invention can accommodate a wide range of pressures such as from 3,000 psi to 10,000 psi, for example. 
     It will be appreciated that the steel liner  24  may be open at both ends, and domed caps  28  can be installed on each end in the same manner as described above. In either case, the domed cap(s)  28  allow access to the piston  18  for repair and/or replacement, thus making the accumulator  10  repairable. 
     As will also be appreciated, an autofrettage process may be performed on the steel liner  24 . After such process, the steel liner bore may be finish machined for accepting the piston  18 . This allows for a close tolerance piston seal and longer fatigue life on the steel liner  24 . 
     As an example, one manner in which an accumulator in accordance with the invention can be made includes starting with a tubular blank such as a steel liner for a steel piston accumulator. The steel blank has a starting wall thickness that is then machined down to decrease the wall thickness thereby reducing weight. At the same time, the radially outwardly extending shoulder is formed at an end of the sleeve surrounding an opening. Although machining is preferably, the shoulder could be formed by other processes, such as forging. The machined liner is then overwrapped with a composite wrap to increase its strength in the hoop direction. The opening of the liner is then closed with a cap as set forth above. This results in a repairable, reduced weight accumulator having pressure capacities similar to the full weight conventional steel piston accumulator. 
     Although the invention has been at least partially described in the context of a hybrid transmission system for a vehicle, the invention is applicable to a wide variety of hydraulic and/or pneumatic systems, and is particularly applicable to mobile systems where reduced vehicle weight can increase efficiency. 
     Although the invention has been shown and described with respect to a certain preferred embodiment or embodiments, it is obvious that equivalent alterations and modifications will occur to others skilled in the art upon the reading and understanding of this specification and the annexed drawings. In particular regard to the various functions performed by the above described elements (components, assemblies, devices, compositions, etc.), the terms (including a reference to a “means”) used to describe such elements are intended to correspond, unless otherwise indicated, to any element which performs the specified function of the described element (i.e., that is functionally equivalent), even though not structurally equivalent to the disclosed structure which performs the function in the herein illustrated exemplary embodiment or embodiments of the invention. In addition, while a particular feature of the invention may have been described above with respect to only one or more of several illustrated embodiments, such feature may be combined with one or more other features of the other embodiments, as may be desired and advantageous for any given or particular application.