Abstract:
The present invention provides the art with a scroll machine which has a plurality of built-in volume ratios along with their respective design pressure ratios. The incorporation of more than one built-in volume ratio allows a single compressor to be optimized for more than one operating condition. The operating envelope for the compressor will determine which of the various built-in volume ratios is going to be selected. Each volume ratio includes a discharge passage extending between one of the pockets of the scroll machine and the discharge chamber. All but the highest volume ration utilize a valve controlling the flow through the discharge passage.

Description:
FIELD OF THE INVENTION 
     The present invention relates generally to scroll machines. More particularly, the present invention relates to a dual volume ratio scroll machine, having a multi-function floating seal system which utilizes flip seals. The scroll machine has the ability to operate at two design pressure ratios. 
     BACKGROUND AND SUMMARY OF THE INVENTION 
     A class of machines exists in the art generally known as scroll machines which are used for the displacement of various types of fluids. Those scroll machines can be configured as an expander, a displacement engine, a pump, a compressor, etc., and the features of the present invention are applicable to any one of these machines. For purposes of illustration, however, the disclosed embodiments are in the form of a hermetic refrigerant compressor. 
     Scroll-type apparatus have been recognized as having distinct advantages. For example, scroll machines have high isentropic and volumetric efficiency, and hence are small and lightweight for a given capacity. They are quieter and more vibration free than many compressors because they do not use large reciprocating parts (e.g. pistons, connecting rods, etc.). All fluid flow is in one direction with simultaneous compression in plural opposed pockets which results in less pressure-created vibrations. Such machines also tend to have high reliability and durability because of the relatively few moving parts utilized, the relatively low velocity of movement between the scrolls, and an inherent forgiveness to fluid contamination. 
     Generally speaking, a scroll apparatus comprises two spiral wraps of similar configuration, each mounted on a separate end plate to define a scroll member. The two scroll members are interfitted together with one of the scroll wraps being rotationally displaced 180 degrees from the other. The apparatus operates by orbiting one scroll member (the orbiting scroll member) with respect to the other scroll member (the non-orbiting scroll) to produce moving line contacts between the flanks of the respective wraps. These moving line contacts create defined moving isolated crescent-shaped pockets of fluid. The spiral scroll wraps are typically formed as involutes of a circle. Ideally, there is no relative rotation between the scroll members during operation, the movement is purely curvilinear translation (no rotation of any line on the body). The relative rotation between the scroll members is typically prohibited by the use of an Oldham coupling. 
     The moving fluid pockets carry the fluid to be handled from a first zone in the scroll machine where a fluid inlet is provided, to a second zone in the scroll machine where a fluid outlet is provided. The volume of the sealed pocket changes as it moves from the first zone to the second zone. At any one instant of time, there will be at least one pair of sealed pockets, and when there are several pairs of sealed pockets at one time, each pair will have different volumes. In a compressor, the second zone is at a higher pressure than the first zone and it is physically located centrally within the machine, the first zone being located at the outer periphery of the machine. 
     Two types of contacts define the fluid pockets formed between the scroll members. First, there is axially extending tangential line contacts between the spiral faces or flanks of the wraps caused by radial forces (“flank sealing”). Second, there are area contacts caused by axial forces between the plane edge surfaces (the “tips”) of each wrap and the opposite end plate (“tip sealing”). For high efficiency, good sealing must be achieved for both types of contacts, however, the present invention is concerned with tip sealing. 
     To maximize efficiency, it is important for the wrap tips of each scroll member to sealingly engage the end plate of the other scroll so that there is minimum leakage therebetween. One way this has been accomplished, other than using tip seals (which are very difficult to assembly and which often present reliability problems) is by using fluid under pressure to axially bias one of the scroll members against the other scroll member. This of course, requires seals in order to isolate the biasing fluid at the desired pressure. Accordingly, there is a continuing need in the field of scroll machines for axial biasing techniques—including improved seals to facilitate the axial biasing. 
     One aspect of the present invention provides the art with a unique sealing system for the axial biasing chamber of a scroll-type apparatus. The seals of the present invention are embodied in a scroll compressor and suited for use in machines which use discharge pressure alone, discharge pressure and an independent intermediate pressure, or solely an intermediate pressure only, in order to provide the necessary axial biasing forces to enhance tip sealing. In addition, the seals of the present invention are suitable particularly for use in applications which bias the non-orbiting scroll member towards the orbiting scroll member. 
     A typical scroll machine which is used as a scroll compressor for an air conditioning application is a single volume ratio device. The volume ratio of the scroll compressor is the ratio of the gas volume trapped at suction closing to the gas volume at the onset of discharge opening. The volume ratio of the typical scroll compressor is “built-in” since it is fixed by the size of the initial suction pocket and the length of the active scroll wrap. The built-in volume ratio and the type of refrigerant being compressed determine the single design pressure ratio for the scroll compressor where compression lossed due to pressure ratio mismatch is avoided. The design pressure ratio is generally chosen to closely match the primary compressor rating point, however, it may be biased towards a secondary rating point. 
     Scroll compressor design specifications for air conditioning applications typically include a requirement that the motor which drives the scroll members must be able to withstand a reduced supply voltage without overheating. While operating at this reduced supply voltage, the compressor must operate at a high-load operating condition. When the motor is sized to meet the reduced supply voltage requirement, the design changes to the motor will generally conflict with the desire to maximize the motor efficiency at the primary compressor rating point. Typically, the increasing of motor output torque will improve the low voltage operation of the motor but this will also reduce the compressor efficiency at the primary rating point. Conversely, any reduction that can be made in the design motor torque while still being able to pass the low-voltage specification allows the selection of a motor which will operate at a higher efficiency at the compressor primary rating point. 
     Another aspect of the present invention improves the operating efficiency of the scroll compressor through the existence of a plurality of built-in volume ratios and their corresponding design pressure ratios. For exemplary purposes, the present invention is described in a compressor having two built-in volume ratios and two corresponding design pressure ratios. It is to be understood that additional built-in volume ratios and corresponding design pressure ratios could be incorporated into the compressor if desired. 
     Other advantages and objects of the present invention will become apparent to those skilled in the art from the subsequent detailed description, appended claims and drawings. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     In the drawings which illustrate the best mode presently contemplated for carrying out the present invention: 
     FIG. 1 is a vertical sectional view of a scroll type refrigerant compressor incorporating the sealing system and the dual volume ratio in accordance with the present invention; 
     FIG. 2 is a cross-sectional view of the refrigerant compressor shown in FIG. 1, the section being taken along line  2 — 2  thereof; 
     FIG. 3 is a partial vertical sectional view of the scroll type refrigerant compressor shown in FIG. 1 illustrating the pressure relief systems incorporated into the compressor; 
     FIG. 4 is a cross-sectional view of the refrigerant compressor shown in FIG. 1, the section being taken along line  2 — 2  thereof with the partition removed; 
     FIG. 5 is a typical compressor operating envelope for an air-conditioning application with the two design pressure ratios being identified; 
     FIG. 6 is an enlarged view of a portion of a compressor in accordance with another embodiment of the present invention; 
     FIG. 7 is an enlarged view of a portion of a compressor in accordance with another embodiment of the present invention; 
     FIG. 8 is an enlarged view of a portion of a compressor in accordance with another embodiment of the present invention; 
     FIG. 9 is an enlarged view of a portion of a compressor in accordance with another embodiment of the present invention; 
     FIG. 10 is an enlarged view of a portion of a compressor in accordance with another embodiment of the present invention; 
     FIG. 11 is an enlarged plan view of a portion of the sealing system according to the present invention shown in FIG. 3; 
     FIG. 12 is an enlarged vertical sectional view of circle  4 — 4  shown in FIG. 2; 
     FIG. 13 is a cross-sectional view of a seal groove in accordance with another embodiment of the present invention; and 
     FIG. 14 is a cross-sectional view of a seal groove in accordance with another embodiment of the present invention. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Although the principles of the present invention may be applied to many different types of scroll machines, they are described herein, for exemplary purposes, embodied in a hermetic scroll compressor, and particularly one which has been found to have specific utility in the compression of refrigerant for air conditioning and refrigeration systems. 
     Referring now to the drawings in which like reference numerals designate like or corresponding parts throughout the several views, there is shown in FIGS. 1 and 2 a scroll compressor incorporating the unique dual volume-ratio in accordance with the present invention which is designated generally by the reference numeral  10 . Scroll compressor  10  comprises a generally cylindrical hermetic shell  12  having welded at the upper end thereof a cap  14  and at the lower end thereof a base  16  having a plurality of mounting feet (not shown) integrally formed therewith. Cap  14  is provided with a refrigerant discharge fitting  18  which may have the usual discharge valve therein (not shown). Other major elements affixed to the shell include a transversely extending partition  22  which is welded about its periphery at the same point that cap  14  is welded to shell  12 , a main bearing housing  24  which is suitably secured to shell  12  and a lower bearing housing  26  having a plurality of radially outwardly extending legs each of which is also suitably secured to shell  12 . A motor stator  28  which is generally square in cross-section but with the corners rounded off is press fitted into shell  12 . The flats between the rounded corners on the stator provide passageways between the stator and shell, which facilitate the return flow of lubricant from the top of the shell to the bottom. 
     A drive shaft or crankshaft  30  having an eccentric crank pin  32  at the upper end thereof is rotatably journaled in a bearing  34  in main bearing housing  24  and a second bearing  36  in lower bearing housing  26 . Crankshaft  30  has at the lower end a relatively large diameter concentric bore  38  which communicates with a radially outwardly inclined smaller diameter bore  40  extending upwardly therefrom to the top of crankshaft  30 . Disposed within bore  38  is a stirrer  42 . The lower portion of the interior shell  12  defines an oil sump  44  which is filled with lubricating oil to a level slightly above the lower end of a rotor  46 , and bore  38  acts as a pump to pump lubricating fluid up the crankshaft  30  and into passageway  40  and ultimately to all of the various portions of the compressor which require lubrication. 
     Crankshaft  30  is rotatively driven by an electric motor including stator  28 , windings  48  passing therethrough and rotor  46  press fitted on crankshaft  30  and having upper and lower counterweights  50  and  52 , respectively. 
     The upper surface of main bearing housing  24  is provided with an annular flat thrust bearing surface  54  on which is disposed an orbiting scroll member  56  having the usual spiral vane or wrap  58  extending upward from an end plate  60 . Projecting downwardly from the lower surface of end plate  60  of orbiting scroll member  56  is a cylindrical hub having a journal bearing  62  therein and in which is rotatively disposed a drive bushing  64  having an inner bore  66  in which crank pin  32  is drivingly disposed. Crank pin  32  has a flat on one surface which drivingly engages a flat surface (not shown) formed in a portion of bore  66  to provide a radially compliant driving arrangement, such as shown in assignee&#39;s U.S. Pat. No. 4,877,382, the disclosure of which is hereby incorporated herein by reference. An Oldham coupling  68  is also provided positioned between orbiting scroll member  56  and bearing housing  24  and keyed to orbiting scroll member  56  and a non-orbiting scroll member  70  to prevent rotational movement of orbiting scroll member  56 . 
     Non-orbiting scroll member  70  is also provided having a wrap  72  extending downwardly from an end plate  74  which is positioned in meshing engagement with wrap  58  of orbiting scroll member  56 . Non-orbiting scroll member  70  has a centrally disposed discharge passage  76  which communicates with an upwardly open recess  78  which in turn is in fluid communication with a discharge muffler chamber  80  defined by cap  14  and partition  22 . A first and a second annular recess  82  and  84  are also formed in non-orbiting scroll member  70 . Recesses  82  and  84  define axial pressure biasing chambers which receive pressurized fluid being compressed by wraps  58  and  72  so as to exert an axial biasing force on non-orbiting scroll member  70  to thereby urge the tips of respective wraps  58 ,  72  into sealing engagement with the opposed end plate surfaces of end plates  74  and  60 , respectively. Outermost recess  82  receives pressurized fluid through a passage  86  and innermost recess  84  receives pressurized fluid through a plurality of passages  88 . Disposed between non-orbiting scroll member  70  and partition  22  are three annular pressure actuated seals  90 ,  92  and  94 . Seals  90  and  92  isolate outermost recess  82  from a suction chamber  96  and innermost recess  84  while seals  92  and  94  isolate innermost recess  84  from outermost recess  82  and discharge chamber  80 . 
     Muffler plate  22  includes a centrally located discharge port  100  which receives compressed refrigerant from recess  78  in non-orbiting scroll member  70 . When compressor  10  is operating at its full capacity or at its highest design pressure ratio, port  100  discharges compressed refrigerant to discharge chamber  80 . Muffler plate  22  also includes a plurality of discharge passages  102  located radially outward from discharge port  100 . Passages  102  are circumferentially spaced at a radial distance where they are located above innermost recess  84 . When compressor  10  is operating at its reduced capacity or at its lower design pressure ratio, passages  102  discharge compressed refrigerant to discharge chamber  80 . The flow of refrigerant through passages  102  is controlled by a valve  104  mounted on partition  22 . A valve stop  106  positions and maintains valve  104  on muffler plate  22  such that it covers and closes passages  102 . 
     Referring now to FIGS. 3 and 4, a temperature protection system  110  and a pressure relief system  112  are illustrated. Temperature protection system  110  comprises an axially extending passage  114 , a radially extending passage  116 , a bi-metallic disc  118  and a retainer  120 . Axial passage  114  intersects with radial passage  116  to connect recess  84  with suction chamber  96 . Bi-metallic disc  118  is located within a circular bore  122  and it includes a centrally located indentation  124  which engages axial passage  114  to close passage  114 . Bi-metallic disc  118  is held in position within bore  122  by retainer  120 . When the temperature of refrigerant in recess  84  exceeds a predetermined temperature, bimetallic disc  118  will snap open or move into a domed shape to space indentation  124  from passage  114 . Refrigerant will then flow from recess  84  through a plurality of holes  126  in disc  118  into passage  114  into passage  116  and into suction chamber  96 . The pressurized gas within recess  82  will vent to recess  84  due to the loss of sealing for annular seal  92 . 
     When the pressurized gas within recess  84  is vented, annular seal  92  will lose sealing because it, like seals  90  and  94 , are energized in part by the pressure differential between adjacent recesses  82  and  84 . The loss of pressurized fluid in recess  84  will thus cause fluid to leak between recess  82  and recess  84 . This will result in the removal of the axial biasing force provided by pressurized fluid within recesses  82  and  84  which will in turn allow separation of the scroll wrap tips with the opposing end plate resulting in a leakage path between discharge chamber  80  and suction chamber  96 . This leakage path will tend to prevent the build up of excessive temperatures within compressor  10 . 
     Pressure relief system  112  comprises an axially extending passage  128 , a radially extending passage  130  and a pressure relief valve assembly  132 . Axial passage  128  intersects with radial passage  130  to connect recess  84  with suction chamber  96 . Pressure relief valve assembly  132  is located within a circular bore  134  located at the outer end of passage  130 . Pressure relief valve assembly  132  is well known in the art and will therefore not be described in detail. When the pressure of refrigerant within recess  84  exceeds a predetermined pressure, pressure relief valve assembly  132  will open to allow fluid flow between recess  84  and suction chamber  96 . The venting of fluid pressure by valve assembly  132  will affect compressor  10  in the same manner described above for temperature protection system  110 . The leakage path which is created by valve assembly  132  will tend to prevent the build-up of excessive pressures within compressor  10 . The response of valve assembly  132  to excessive discharge pressures is improved if the compressed pocket that is in communication with recess  84  is exposed to discharge pressure for a portion of the crank cycle. This is the case if the length of the active scroll wraps  58  and  72  needed to compress between an upper design pressure ratio  140  and a lower design pressure  142  (FIG. 5) is less then 360°. 
     Referring now to FIG. 5, a typical compressor operating envelope for an air conditioning application is illustrated. Also shown are the relative locations for upper design pressure ratio  140  and lower design pressure ratio  142 . Upper design pressure ratio  140  is chosen to optimize operation of compressor  10  at the motor low-voltage test point. When compressor  10  is operating at this point, the refrigerant being compressed by scroll members  56  and  70  enter discharge chamber  80  through discharge passage  76 , recess  78  and discharge port  100 . Discharge passages  102  are closed by valve  104  which is urged against partition  22  by the fluid pressure within discharge chamber  80 . Increasing the overall efficiency of compressor  10  at design pressure ratio  140  allows the design motor torque to be reduced which yields increased motor efficiency at the rating point. Lower design pressure ratio  142  is chosen to match the rating point for compressor  10  to further improve efficiency. 
     Thus, if the operating point for compressor  10  is above lower design pressure ratio  142 , the gas within the scroll pockets is compressed along the full length of wraps  58  and  72  in the normal manner to be discharged through passage  76 , recess  78  and port  100 . If the operating point for compressor  10  is at or below lower design pressure ratio  142 , the gas within the scroll pockets is able to discharge through passages  102  by opening valve  104  before reaching the inner ends of scroll wraps  58  and  72 . This early discharging of the gas avoids losses due to compression ratio mismatch. 
     Outermost recess  82  acts in a typical manner to offset a portion of the gas separating forces in the scroll compression pockets. The fluid pressure within recess  82  axially bias the vane tips of non-orbiting scroll member  70  into contact with end plate  60  of orbiting scroll member  56  and the vane tips of orbiting scroll member  56  into contact with end plate  74  of non-orbiting scroll member  70 . Innermost recess  84  acts in this typical manner at a reduced pressure when the operating condition of compressor  10  is below lower design pressure ratio  142  and at an increased pressure when the operating condition of compressor  10  is at or above lower design pressure ratio  142 . In this mode, recess  84  can be used to improve the axial pressure balancing scheme since it provides an additional opportunity to minimize the tip contact force. 
     In order to minimize the re-expansion losses created by axial passages  88  and  102  used for early discharge end, the volume defined by innermost recess  84  should be held to a minimum. An alternative to this would be to incorporate a baffle plate  150  into recess  84  as shown in FIGS. 1 and 6. Baffle plate  150  controls the volume of gas that passes into recess  84  from the compression pockets. Baffle plate  150  operates similar to the way that valve plate  104  operates. Baffle plate  150  is constrained from angular motion but it is capable of axial motion within recess  84 . When baffle plate  150  is at the bottom of recess  84  in contact with non-orbiting scroll member  70 , the flow of gas into recess  84  is minimized. Only a very small bleed hole  152  connects the compression pocket with recess  84 . Bleed hole  152  is in line with one of the axial passages  88 . Thus, expansion losses are minimized. When baffle plate  150  is spaced from the bottom of recess  84 , sufficient gas flow for early discharging flows through a plurality of holes  154  offset in baffle plate  150 . Each of the plurality of holes  154  is in line with a respective passage  102  and not in line with any of passages  88 . When using baffle plate  150  and optimizing the response of pressure relief valve assembly  132  by having an active scroll length of 360° between ratios  140  and  142  as described above, the trade off for this increased response will be the possibility of the opening of baffle plate  150 . 
     Referring now to FIG. 6, an enlarged section of recesses  78  and  84  of non-orbiting scroll member  70  is illustrated according to another embodiment of the present invention. In this embodiment, a discharge valve  160  is located within recess  78 . Discharge valve  160  includes a valve seat  162 , a valve plate  164  and a retainer  166 . 
     Referring now to FIG. 7, an enlarged section of recesses  78  and  84  of non-orbiting scroll member  70  is illustrated according to another embodiment of the present invention. In this embodiment valve  104  and baffle plate  150  are connected by a plurality of connecting members  170 . Connecting members  170  require that valve  104  and baffle plate  150  move together. The benefit to connecting valve  104  and baffle plate  150  is to avoid any dynamic interaction between the two. 
     Referring now to FIG. 8, an enlarged section of recesses  78  and  84  of non-orbiting scroll member  70  is illustrated according to another embodiment of the present invention. In this embodiment valve  104  and baffle plate  150  are replaced with a single unitary valve  104 ′. Using single unitary valve  104 ′ has the same advantages as those described for FIG. 7 in that dynamic interaction is avoided. 
     Referring now to FIG. 9, an enlarged section of recesses  78  and  84  of a non-orbiting scroll member  270  is illustrated according to another embodiment of the present invention. Scroll member  270  is identical to scroll member  70  except that a pair of radial passages  302  replace the plurality of passages  102  through partition  22 . In addition, a curved flexible valve  304  located along the perimeter of recess  78  replaces valve  104 . Curved flexible valve  304  is a flexible cylinder which is designed to flex and thus to open radial passages  302  in a similar manner with the way that valve  104  opens passages  102 . The advantage to this design is that a standard partition  22  which does not include passages  102  can be utilized. While this embodiment discloses radial passage  302  and flexible valve  304 , it is within the scope of the present invention to eliminate passage  302  and valve  304  and design annular seal  94  to function the valve between innermost recess  84  and discharge chamber  80 . Since annular seal  94  is a pressure actuated seal, the higher pressure within discharge chamber  80  over the pressure within recess  84  actuates seal  94 . Thus, if the pressure within recess  84  would exceed the pressure within discharge chamber  80 , seal  94  could be designed to open and allow the passage of the high pressure gas. 
     Referring now to FIG. 10, an enlarged section of recess  78  and  84  of a non-orbiting scroll member  370  is illustrated according to another embodiment of the present invention. Scroll member  370  is identical to scroll member  70  except that the pair of radial passages  402  replace the plurality of passages  102  through partition  22 . In addition, a valve  404  is biased against passages  402  by a retaining spring  406 . A valve guide  408  controls the movement of valves  404 . Valves  404  are designed to open radial passages  402  in a similar manner with the way that valve  104  opens passages  102 . The advantage to this design is again that a standard partition  22  which does not include passages  102  can be utilized. 
     While not specifically illustrated, it is within the scope of the present invention to configure each of valves  404  such that they perform the function of both opening passages  402  and minimize the re-expansion losses created through passages  88  in a manner equivalent to that of baffle plate  150 . 
     With reference to FIGS. 1,  2 ,  11  and  12 , annular seals  90 ,  92  and  94  are each configured as an annular L-shaped seal. Outer L-shaped seal  90  is disposed within a groove  200  located within non-orbiting scroll member  70 . One leg of seal  90  extends into groove  200  while the other leg extends generally horizontal, as shown in FIGS. 1,  2  and  12  to provide sealing between non-orbiting scroll member  70  and muffler plate  22 . Seal  90  functions to isolate the bottom of recess  82  from the suction area of compressor  10 . The initial forming diameter of L-shaped seal  90  is less than the diameter of groove  200  such that the assembly of seal  90  into groove  200  requires stretching of seal  90 . Preferably, seal  90  is manufactured from a Teflon® material containing 10% glass when interfacing with steel components. 
     Center L-shaped seal  92  is disposed within a groove  204  located within non-orbiting scroll member  70 . One leg of seal  92  extends into groove  204  while the other leg extends generally horizontal, as shown in FIGS. 1,  2  and  12  to provide sealing between non-orbiting scroll member  70  and muffler plate  22 . Seal  92  functions to isolate the bottom of recess  82  from the bottom of recess  84 . The initial forming diameter of L-shaped seal  92  is less than the diameter of groove  204  such that the assembly of seal  92  into groove  204  requires stretching of seal  92 . Preferably, seal  92  is manufactured from a Teflon® material containing 10% glass when interfacing with steel components. 
     Inner L-shaped seal  94  is disposed within a groove  208  located within non-orbiting scroll member  70 . One leg of seal  94  extends into groove  208  while the other leg extends generally horizontal, as shown in FIGS. 1,  2  and  12  to provide sealing between non-orbiting scroll member  70  and muffler plate  22 . Seal  94  functions to isolate the bottom of recess  84  from the discharge area of compressor  10 . The initial forming diameter area of L-shaped seal  94  is less than the diameter of groove  208  such that the assembly of seal  94  into groove  208  requires stretching of seal  94 . Preferably, seal  94  is manufactured from a Teflon® material containing 10% glass when interfacing with steel components. 
     Seals  90 ,  92  and  94  therefore provide three distinct seals; namely, an inside diameter seal of seal  94 , an outside diameter seal of seal  90 , and a middle diameter seal of seal  92 . The sealing between muffler plate  22  and seal  94  isolates fluid under intermediate pressure in the bottom of recess  84  from fluid under discharge pressure. The sealing between muffler plate  22  and seal  90  isolates fluid under intermediate pressure in the bottom of recess  82  from fluid under suction pressure. The sealing between muffler plate  22  and seal  92  isolates fluid under intermediate pressure in the bottom of recess  84  from fluid under a different intermediate pressure in the bottom of recess  82 . Seals  90 ,  92  and  94  are pressure activated seals as described below. 
     Grooves  200 ,  204  and  208  are all similar in shape. Groove  200  will be described below. It is to be understood that grooves  204  and  208  include the same features as groove  200 . Groove  200  includes a generally vertical outer wall  240 , a generally vertical inner wall  242  and an undercut portion  244 . The distance between walls  240  and  242 , the width of groove  200 , is designed to be slightly larger than the width of seal  90 . The purpose for this is to allow pressurized fluid from recess  82  into the area between seal  90  and wall  242 . The pressurized fluid within this area will react against seal  90  forcing it against wall  240  thus enhancing the sealing characteristics between wall  240  and seal  90 . Undercut  244  is positioned to lie underneath the generally horizontal portion of seal  90  as shown in FIG.  12 . The purpose for undercut  244  is to allow pressurized fluid within recess  82  to act against the horizontal portion of seal  92  urging it against muffler plate  22  to enhance its sealing characteristics. Thus, the pressurized fluid within recess  82  reacts against the inner surface of seal  90  to pressure activate seal  90 . As stated above, grooves  204  and  208  are the same as groove  200  and therefore provide the same pressure activation for seals  92  and  94 . 
     The stretching of seals  90 ,  92  and  94  in order to assemble them into grooves  200 ,  204  and  208 , respectively, aids in keeping the seals within the grooves during operation of compressor  10 . This is important for two reasons. First, the seals must be kept free floating in the grooves in order to minimize the movement of the seal against muffler plate  22 . The movement of the seal is minimized due to the fact that the movement of non-orbiting scroll  70  is accommodated by the movement of seals  90 ,  92  and  94 . Second, it is important that seal  94  seal in only one direction. Seal  94  is used to relieve high intermediate pressure from the bottom of recess  84  during flooded starts. The relieving of this high intermediate pressure reduces inner-scroll pressures and the resultant stress and noise. 
     The unique L-shaped seals  90 ,  92  and  94  of the present invention are relatively simple in construction, easy to install and inspect, and effectively provide the complex sealing functions desired. The unique sealing system of the present invention comprises three L-shaped seals  90 ,  92  and  94  that are “stretched” into place and then pressure activated. The unique seal assembly of the present invention reduces overall manufacturing costs for the compressor, reduces the number of components for the seal assembly, improves durability by minimizing seal wear and provides room to increase the discharge muffler volume for improved damping of discharging pulse without increasing the overall size of the compressor. 
     The seals of the present invention also provide a degree of relief during flooded starts. Seals  90 ,  92  and  94  are designed to seal in only one direction. These seals can then be used to relieve high pressure fluid from the intermediate chambers or recesses  82  and  84  to the discharge chamber during flooded starts, thus reducing inter-scroll pressures and the resultant stress and noise. 
     Referring now to FIG. 13, a groove  300  in accordance with another embodiment of the present invention is illustrated. Groove  300  includes an outwardly angled outer wall  340 , generally vertical inner wall  242  and undercut portion  244 . Thus, groove  300  is the same as groove  200  except that the outwardly angled outer wall  340  replaces generally vertical outer wall  240 . The function, operation and advantages of groove  300  and seal  90  are the same as groove  200  and seal  90  detailed above. The angling of the outer wall enhances the ability of the pressurized fluid within recess  82  to react against the inner surface of seal  90  to pressure activate seal  90 . It is to be understood that grooves  200 ,  204  and  208  can each be configured the same as groove  300 . 
     Referring now to FIG. 14, a seal groove  400  in accordance with another embodiment of the present invention is illustrated. Groove  400  includes outwardly angled outer wall  340  and a generally vertical inner wall  442 . Thus, groove  400  is the same as groove  300  except that undercut portion  244  has been removed. The function, operation and advantages of groove  300  and seal  90  are the same as grooves  200  and  300  and seal  90  as detailed above. The elimination of undercut portion  244  is made possible by the incorporation of a wave spring  450  underneath seal  90 . Wave spring  450  biases the horizontal portion of seal  90  upward toward muffler plate  22  to provide a passage for the pressurized gas within recess  82  to react against the inner surface of seal  90  to pressure activate seal  90 . It is to be understood that grooves  200 ,  204  and  208  can each be configured the same as groove  400 . 
     While the above detailed description describes the preferred embodiment of the present invention, it should be understood that the present invention is susceptible to modification, variation and alteration without deviating from the scope and fair meaning of the subjoined claims.