Patent Publication Number: US-2023145160-A1

Title: High-throughput diaphragm compressor

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims the benefit under 35 U.S.C. § 119(e) of the earlier filing date of U.S. Provisional Patent Application No. 63/277,125 filed on Nov. 8, 2021, the disclosure of which is incorporated by reference herein. 
    
    
     FIELD OF THE INVENTION 
     The present invention is directed to diaphragm compressors and modifications for improving reliability and hydraulic efficiency in high-pressure and/or high-throughput applications. 
     BACKGROUND OF THE INVENTION 
     A diaphragm compressor actuates a diaphragm at high speed to pressurize a process gas. Although some modern applications require process gas at high pressures and/or in large tanks, a conventional diaphragm compressor system is limited by physical constraints, for example the compressor head volume, speed of operation, actuation force, material strength, and the like. 
     SUMMARY OF THE INVENTION 
     A feature and benefit of embodiments is a diaphragm compressor system, comprising: 
     1. A diaphragm compressor system, comprising: 
     a first compressor head comprising: 
     a head cavity, and 
     a diaphragm mounted in the head cavity and dividing the head cavity into a work oil region and a process gas region, 
     the diaphragm configured to actuate from a first position to a second position during a discharge cycle to pressurize process gas in the process gas region from an inlet pressure to a discharge pressure, and discharge the pressurized process gas through the first compressor head; 
     a hydraulic drive comprising a high-pressure circuit of work oil and a low-pressure circuit of work oil, the hydraulic drive configured to pressurize work oil and distribute work oil throughout the diaphragm compressor system provide the pressurized work oil to the first compressor head, the hydraulic drive comprising: 
     a drive housing comprising: 
     a drive cavity, 
     first and second ports, wherein the hydraulic drive the hydraulic drive being configured to provide a variable-pressure supply of work oil to the drive cavity through the first and second ports, and 
     an actuator piston located in the drive cavity, the actuator piston dividing the drive cavity into a first actuation volume oriented toward the first compressor head and in communication with the first port and a second actuation volume oriented away from the first compressor head and in communication with the second port, the actuator piston comprising a first side oriented toward the first actuation volume and a second side oriented toward the second actuation volume, the actuator piston being configured to intensify work oil in the work oil region of the first compressor head by compressing work oil in the first actuation volume, 
     wherein, after a discharge cycle of the first compressor head: 
     the hydraulic drive is configured to provide work oil from the high-pressure circuit through both the first and second ports to the first and second actuation volumes to substantially balance the pressure on the first and second sides of the actuator piston , and 
     the system is configured to supply the first compressor head with process gas to begin a supply cycle, the supply of process gas driving the diaphragm toward its first position, intensifying the work oil in the work oil region, increasing pressure in the first actuation volume, and thereby actuating the actuator piston to move away from the first compressor head. 
     2. The diaphragm compressor system of claim  1 , wherein, as the actuator piston is actuated in response to the supply of process gas, work oil exits from the second actuation volume and thereafter enters the first actuation volume. 
     3. The diaphragm compressor system of claim  1 , wherein, as the actuator piston is actuated in response to the supply of process gas, work oil exits from the second actuation volume to the high-pressure circuit, and work oil enters the first actuation volume from the high-pressure circuit. 
     4. A diaphragm compressor system, comprising: 
     a first compressor head comprising: 
     a head cavity, and 
     a diaphragm mounted in the head cavity and dividing the head cavity into a work oil region and a process gas region, 
     the diaphragm configured to actuate from a first position to a second position during a discharge cycle to pressurize process gas in the process gas region from an inlet pressure to a discharge pressure, and discharge the pressurized process gas through the first compressor head; 
     a hydraulic drive comprising a high-pressure circuit of work oil, a medium-pressure circuit of work oil, and a low-pressure circuit of work oil, the hydraulic drive configured to pressurize work oil and distribute work oil throughout the diaphragm compressor system, the hydraulic drive comprising: 
     a drive housing comprising: 
     a drive cavity, 
     first and second ports, the hydraulic drive being configured to provide a variable-pressure supply of work oil to the drive cavity through the first and second ports, and 
     an actuator piston located in the drive cavity, the actuator piston dividing the drive cavity into a first actuation volume oriented toward the first compressor head and in communication with the first port and a second actuation volume oriented away from the first compressor head and in communication with the second port, the actuator piston comprising a first side oriented toward the first actuation volume and a second side oriented toward the second actuation volume, the actuator piston being configured to intensify work oil in the work oil region of the first compressor head by compressing work oil in the first actuation volume, 
     wherein, after a discharge cycle of the first compressor head: 
     the hydraulic drive is configured to close off work oil from the high-pressure circuit to the second actuation volume, and provide work oil from the medium-pressure circuit through both the first and second ports to the first and second actuation volumes to substantially balance the pressure on the first and second sides of the actuator piston at an intermediate pressure between the pressures of the medium-pressure circuit and the high-pressure circuit, and 
     the system is configured to supply the first compressor head with process gas to begin a supply cycle, the supply of process gas driving the diaphragm toward its first position, intensifying the work oil in the work oil region, increasing pressure in the first actuation volume, and thereby actuating the actuator piston to move away from the first compressor head. 
     5. The diaphragm compressor system of claim  4 , wherein, as the actuator piston is actuated in response to the supply of process gas, work oil exits from the second actuation volume and thereafter enters the first actuation volume. 
     6. The diaphragm compressor system of claim  4 , the hydraulic drive further comprising a medium-pressure circuit comprising a MP-rail connector, and a high-pressure circuit comprising a HP-rail connector, 
     wherein, after force coupling movement of the actuator piston, the medium-pressure circuit drives movement of the actuator piston while the second actuation volume is opened to the low-pressure circuit, and subsequently the high-pressure circuit drives movement of the actuator piston. 
     7. The diaphragm compressor system of claim  6 , wherein the first actuation volume remains open to the medium-pressure circuit while the high-pressure circuit drives movement of the actuator piston, and high-pressure oil flows through the MP-rail connector. 
     8. A diaphragm compressor system, comprising: 
     first and second compressor heads each comprising: 
     a diaphragm mounted in a head cavity and dividing the head cavity into a work oil region and a process gas region, 
     the diaphragm configured to: 
     during a discharge cycle, actuate from a first position to a second position in response to intensified work oil in the respective work oil region and thereby pressurizing process gas in the process gas region from an inlet pressure to a discharge pressure, and 
     during a supply cycle, move from the second position to the first position in response to a supply of process gas at the inlet pressure into the process gas region; and 
     a hydraulic drive comprising: 
     a plurality of pressure circuits of work oil, 
     a hydraulically-driven actuator piston configured to drive alternatingly in first and second stroke directions, the first stroke direction intensifying work oil in the work oil region of the first compressor head and the second stroke direction intensifying work oil in the work oil of the second compressor head, 
     a drive cavity for the actuator piston, the drive cavity comprising a first actuation volume on a first side of the actuator piston and a second actuation volume on a second side of the actuator piston, and 
     one or more valves configured to selectively open or close one or more of the plurality of pressure circuits to one or more of the first and second actuation volumes, 
     wherein the hydraulic drive is configured to be force coupled with each of the first and second compressor heads such that, during a supply cycle of the first compressor head, the movement of the respective diaphragm toward its first position is configured to drive the actuator piston in the second stroke direction, and 
     wherein, upon completion of a discharge cycle in the first compressor head and corresponding completion of driving the actuator piston in the first stroke direction, the one or more valves are configured to balance pressures in the first and second actuation volumes by opening both volumes to a predetermined one of the plurality of pressure circuits. 
     9. The diaphragm compressor system of claim  8 , wherein, as the actuator piston is actuated in response to the supply of process gas, work oil exits from the second actuation volume to the high-pressure circuit, and work oil enters the first actuation volume from the high-pressure circuit. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The invention can be completely understood in consideration of the following detailed description of various embodiments of the invention in connection with the accompanying drawings, in which: 
         FIG.  1    is a front perspective view of a high-throughput compressor system with stacked compressor modules in accord with embodiments of the present disclosure. 
         FIG.  2    is a top front perspective view of an embodiment of a compressor module of the system of  FIG.  1    in accord with embodiments of the present disclosure. 
         FIG.  3    is a bottom rear perspective view of the compressor module of  FIG.  2   . 
         FIG.  4    is front elevation view of the compressor module of  FIG.  2   . 
         FIG.  5    is side elevation view of the compressor module of  FIG.  2   . 
         FIG.  6    is top sectional view of the compressor module of  FIG.  2   . 
         FIG.  7    is side sectional view of the compressor module of  FIG.  2   . 
         FIG.  8    is an enlarged partial view of  FIG.  7   . 
         FIG.  9    is a side sectional view of another embodiment of a compressor module of the system of  FIG.  1    in accord with embodiments of the present disclosure. 
         FIG.  10    is a side sectional view of still another embodiment of a compressor module of the system of  FIG.  1    in accord with embodiments of the present disclosure. 
         FIG.  11    is an enlarged partial view of  FIG.  10   . 
         FIG.  12    is a sectional view of a compressor head for a compressor module in accord with embodiments of the present disclosure. 
         FIG.  13    is a top perspective wireframe view of another compressor head for a compressor module in accord with embodiments of the present disclosure. 
         FIG.  14    is a schematic view of a hydraulically-driven compressor module with two compressor heads force coupled in accord with embodiments of the present disclosure. 
         FIG.  15    is a hydraulic circuit diagram of a hydraulically-driven compressor module with three pressure rails in accord with embodiments of the present disclosure. 
         FIG.  16    is a schematic view of a hydraulically-driven compressor module with three pressure rails in accord with embodiments of the present disclosure. 
         FIG.  17    is a hydraulic circuit diagram of a hydraulically-driven compressor module with three pressure rails in accord with embodiments of the present disclosure. 
         FIG.  18 A  is a cross-sectional view of a main stage valve of the high-throughput compressor system in accord with embodiments of the present disclosure in a vent position. 
         FIG.  18 B  is a cross-sectional view of the main stage valve of  FIG.  18 A  in a supply position. 
         FIG.  19    is a partial cross-sectional view of a compressor module of the system of  FIG.  1    in accord with embodiments of the present disclosure. 
         FIG.  20    is a top perspective wireframe view of a valve manifold of the compressor module of  FIG.  2    in accord with embodiments of the present disclosure. 
         FIG.  21    is a top sectional view of a hydraulic clamp actuator of the system of  FIG.  1    in accord with embodiments of the present disclosure. 
         FIG.  22    is a hydraulic circuit diagram of a staged arrangement of multiple stacks of the high-throughput compressor system of  FIG.  1    in accord with embodiments of the present disclosure. 
         FIG.  23    is a hydraulic circuit diagram of another staged arrangement of multiple stacks of the high-throughput compressor system of  FIG.  1    in accord with embodiments of the present disclosure. 
         FIG.  24    is a front perspective view of another high-throughput compressor system with stacked compressor modules in accord with embodiments of the present disclosure. 
         FIG.  25    is a front perspective view of still another high-throughput compressor system with stacked compressor modules in accord with embodiments of the present disclosure. 
         FIG.  26    is a front perspective view of yet another high-throughput compressor system with stacked compressor modules in accord with embodiments of the present disclosure. 
         FIG.  27    is a cross-sectional view of the system of  FIG.  26   . 
         FIG.  28    is a front perspective view of another high-throughput compressor system with stacked compressor modules in accord with embodiments of the present disclosure. 
         FIG.  29    is a front perspective view of still another high-throughput compressor system with stacked compressor modules in accord with embodiments of the present disclosure. 
         FIG.  30    is a schematic view of multiple hydraulically-driven compressor modules with a common intensifier in accord with embodiments of the present disclosure. 
         FIG.  31    is a schematic view of multiple hydraulically-driven compressor modules with a common control valve in accord with embodiments of the present disclosure. 
         FIG.  32    is a schematic view of a hydraulically-driven compressor system with direct hydraulic actuation in accord with embodiments of the present disclosure. 
         FIG.  33    is a schematic view of a hydraulically-driven compressor module with an active oil injection system in accord with embodiments of the present disclosure. 
         FIG.  34    is a hydraulic circuit diagram of a hydraulically-driven compressor module with force coupling in a high-pressure recovery arrangement. 
         FIG.  35    is a diagram of valve phases and piston positions throughout piston stroke cycles of the arrangement of  FIG.  34   . 
         FIGS.  36 A- 36 E  are simplified diagrams of sequential pressure states throughout piston stroke cycles of the arrangement of  FIG.  34   . 
         FIG.  37    is a hydraulic circuit diagram of a hydraulically-driven compressor module with force coupling in a medium-pressure shuffling arrangement. 
         FIG.  38    is a diagram of valve phases and piston positions throughout piston stroke cycles of the arrangement of  FIG.  37   . 
         FIGS.  39 A- 39 F  are simplified diagrams of sequential pressure states throughout piston stroke cycles of the arrangement of  FIG.  37   . 
     
    
    
     While the invention is amenable to various modifications and alternative forms, specifics thereof have been depicted by way of example in the drawings and will be described in detail. It should be understood, however, that the intention is not to limit the invention to the particular embodiments described. On the contrary, the intention is to cover all modifications, equivalents, and alternatives falling within the spirit and scope of the invention as defined by the appended claims. 
     DETAILED DESCRIPTION 
     As shown in  FIG.  1   , in embodiments of the present disclosure, a high-throughput compressor system  200  comprises multiple compressor modules  100 , for example compressor modules  100 A,  100 B,  100 C,  100 D, collectively referred to as a stack  201  of compressor modules  100 . Each compressor module  100  is a diaphragm compressor with one or more compressor heads  31 ,  51  each having a diaphragm  5 . 
     The process gas may be any gas suitable for pressurization for any use. In embodiments, the process gas is hydrogen. For embodiments designed for filling stations for hydrogen fuel cell vehicles, the required outlet pressure of the high-throughput compressor system  200  may be approximately 10,000-12,000 psi. In embodiments, the target pressure of stored hydrogen in a tank (e.g., at a vehicle filling station) is up to about 14,500 psi to account for pressure losses in, e.g., storage and transfer. Therefore, the corresponding discharge pressure of the process gas from the high-throughput compressor system  200  in such embodiments is about 15,000 psi. 
     Diaphragm Compressor 
     Applicable embodiments of the architecture and function of an individual diaphragm compressor  1  are shown in  FIGS.  12  and  13    and may be similar to the compressor disclosed in U.S. patent application Ser. No. 17/522,896, the entire contents of which are incorporated herein by reference and for all purposes. Relative to the present disclosure, the compressor  1  in U.S. Ser. No. 17/522,896 constitutes an embodiment of each diaphragm compressor head  31 ,  51 , for example the diaphragm compressor heads of the compressor module  100 . Similar diaphragm compressors and related systems are also disclosed in U.S. Provisional Application Nos. 63/111,356 filed Nov. 9, 2020 and 63/277,125 filed on Nov. 8, 2021, and U.S. patent application Ser. No. 17/522,896 filed Nov. 9, 2021, the entire contents of which are incorporated herein by reference and for all purposes. 
     In embodiments, the diaphragm compressor  1  is driven by a diaphragm piston  3  (also referred to as a high-pressure oil piston) that moves a volume of work oil (i.e., hydraulic fluid) through the compressor  1  suction and discharge cycles. Process gas compression occurs as the volume of work oil is pushed towards the diaphragm  5  by a diaphragm piston  3  to fill a work oil region  35  in a work oil head support plate  8  (or lower plate or oil plate), exerting a uniform force against the bottom of the diaphragm  5 . This deflects the diaphragm  5  into an upper cavity in a gas plate  6  that is filled with the process gas, also referred to as a process gas region  36 . The deflection of the diaphragm  5  against the upper cavity of gas plate  6  first compresses the process gas and then expels it through an outlet port  9  comprising a discharge check valve. As the oil piston  3  reverses to begin the suction cycle, the diaphragm  5  is drawn downward towards the oil plate  8  while the inlet check valve at the inlet port  7  opens and fills the process gas region  36  with a fresh charge of process gas at an inlet pressure. The diaphragm piston  3  reaches the end of its stroke before beginning its next stroke, and the compression cycle is repeated. 
     In embodiments, the compressor head  31  comprises a process gas head support plate  6 , a work oil head support plate  8 , and a diaphragm  5 . The process gas head support plate  6  comprises a process gas inlet port  7  operatively connected to an inlet check valve and a process gas outlet port  9  operatively connected to a discharge check valve. In certain embodiments, the work oil head support plate  8  comprises an inlet  33  operatively connected to one or more inlet check valves  45 , and an outlet  34  operatively connected to one or more relief valves  42  (inlet check valves and relief valves shown schematically in  FIG.  33   ). A head cavity  15  is defined between the process gas head support plate  6  and the work oil head support plate  8 . In certain embodiments, the compressor head  31  comprises a piston bore  32  extending toward the work oil head support plate  8  and sized to receive the diaphragm piston  3 . In other embodiments, there is no piston bore  32  and the diaphragm piston  3  is configured to remain substantially within the drive housing  114 . 
     The diaphragm  5  is mounted in the head cavity  15  between the process gas head support plate  6  and the work oil head support plate  8  and divides the head cavity into a work oil region  35  and a process gas region  36 . The diaphragm piston  3  defines the volume of the work oil region  35  between a top face of the diaphragm piston  3  and a bottom face of the diaphragm  5 . Because the diaphragm piston  3  and diaphragm  5  are dynamic, the volume of the work oil region  35  is variable. 
     The diaphragm  5  is configured to actuate from a first position proximate the work oil head support plate  8  (e.g., in contact with the work oil head support plate or fully extended toward the work oil head support plate) to a second position proximate the process gas head support plate  6  during a discharge cycle to pressurize process gas in the process gas region  36  from an inlet pressure to a discharge pressure, and discharge the pressurized process gas through the outlet port  9 . During a suction cycle of the compressor head  31 , the diaphragm  5  is configured to move from the second position to the first position to fill the process gas region  36  with process gas at the inlet pressure. In embodiments, the diaphragm  5  is a diaphragm set comprising a plurality of diaphragm plates sandwiched together and acting in unison, for example two, three, four, or more diaphragm plates may comprise a diaphragm set. In certain embodiments, the diaphragm plates are made from a metal. In other embodiments, the diaphragm plates are made from different metals. In other embodiments, one or more of the diaphragm plates are not made from metal. In certain embodiments, the diaphragm  5  includes three plates, the three plates comprising stainless steel in the outside plates, and brass on the inside plate. 
     Compressor heads applicable to embodiments of the present disclosure may be provided in any of various sizes and compression ratios. In embodiments, an individual compressor head  31  may be configured for a pressure range of process gas outlet of 200 psi to 15,000 psi. In other embodiments, a compressor head  31  may be configured for a maximum pressure range of 40 psi to 30,000 psi. In still further embodiments, a compressor head  31  may be configured for a pressure range of 300 psi to 45,000 psi. In certain embodiments, the aforementioned compressor heads  31  may be run at pressures below 200 psi, 40 psi, and 300 psi, respectively. In some embodiments, a compressor head  31  can have a compression ratio range of 0.25:1, 1:1, 2:1, 3:1, 4:1, 5:1, 6:1, 10:1, 20:1, and ranges therebetween. 
     Hydraulically-Driven Compressor Modules 
     Referring to  FIGS.  2 - 8   , an embodiment of a compressor module  100  is shown. Applicable embodiments of the architecture and function of the individual compressor module  100  are discussed in U.S. patent application Ser. No. 17/522,896 (therein referred to as a “compressor system”). The compressor module  100  comprises a first compressor head  31  and a second compressor head  51 . The compressor module  100  in some embodiments is hydraulically driven by a hydraulic drive  110  that is configured to intensify or pressurize work oil and provide the intensified work oil to the first and second compressor heads  31 ,  51 . In embodiments, the hydraulic drive  110  comprises an actuator  112 , a drive housing  114  defining a drive cavity  116 , and a hydraulic power unit  118  (“HPU”) providing pressurized hydraulic fluid at a pressure, which effectively supplies a pressurized circuit  120  (also referred to broadly as a pressure rail, volume of work oil at a given pressure, or flow of work oil at a given pressure). In embodiments, the hydraulic drive  110  includes one or more pressurized circuits  120  provided by one or more HPUs  118 , and in further embodiments, the actuator  112  comprises a piston subassembly  122 . In some embodiments, the hydraulic drive  110  is configured to provide a variable-pressure supply of work oil to the drive cavity  116  from one or more of: different pressures of work oil in a one or more pressurized circuits  120 , variable areas of components of the piston subassembly  122  (e.g., a variable-area architecture), and/or variable control of the piston subassembly. 
     In certain embodiments, the piston subassembly  122  (e.g. as shown in  FIG.  6   ) comprises the diaphragm piston  3  mounted at least partially in the drive housing  114  and extending into the piston bore  32  ( FIG.  6   , see also  FIG.  12   ). In some embodiments, the piston bore  32  is formed partially or completely in the drive housing  114  (e.g.,  FIG.  6   ). A first variable volume region  54  comprises the work oil region  35  of the compressor head  31  along with the available volume of the piston bore  32 ; in other words, the first variable volume region  54  is defined between the diaphragm piston  3  and the diaphragm  5  of the corresponding compressor head  31 . The piston subassembly  122  comprises an actuator piston  126  located in the drive cavity  116  and coupled (directly or indirectly, for example rigidly coupled, mechanically linked, or hydraulically coupled) to the diaphragm piston  3 , the actuator piston defining an actuator piston axis  208 . The diaphragm piston  3  is coupled to the actuator piston  126  to move in response to movement of the actuator piston  126 . In some embodiments, the diaphragm piston  3  is mechanically rigidly fixed to the actuator piston  126  or, as shown in  FIGS.  6 - 10   , formed as a unitary one-piece part with the actuator piston. In other words, the diaphragm piston  3  may be one control area and the actuator piston  126  another control area of the same unitary piston; likewise, a second diaphragm piston  140  (discussed below) may be a third control area. 
       FIGS.  2 - 8 ,  9 , and  10 - 11    illustrate embodiments of a compressor module  100  applicable to the present disclosure that is dual-headed and comprises the compressor head  31  and the second compressor head  51 .  FIGS.  14 - 17    schematically illustrate embodiments of the hydraulic drive  110  for a dual-headed compressor module  100 , although the hydraulic drive is applicable to a compressor module having any number of heads, for example 1-6 heads. The second compressor head  51  is actuated by a second diaphragm piston  140  defining a second variable volume region  142 . In some embodiments, the piston subassembly  122  is mounted in the drive cavity  116  of the drive housing  114  and a plurality of variable volumes are provided between the piston subassembly  122  and the drive housing  114 . 
     As shown in  FIG.  8   , a first actuation volume  144  is defined on the side of the actuator piston  126  proximate the compressor head  31 , and a second actuation volume  146  is defined on the opposite side of the actuator piston and proximate the second compressor head  51 . Other embodiments may include one, three, or more variable volumes. Due to movement of the piston subassembly  122 , the first and second actuation volumes  144 ,  146  are variable in volume and defined as the volume between the respective first and second sides  143 ,  145  of the actuator piston  126  and the interior of the respective first and second diaphragm pistons  3 ,  140 . The variable volume is a result of the movement of the actuator piston  126  back and forth. As discussed below, in certain operating states, the first and second actuation volumes  144 ,  146  also serve a damping function against the actuator piston  126  as it is being driven. 
     Referring to  FIGS.  6 - 11   , the drive housing  114  also comprises a plurality of ports  147  in communication with the first and second actuation volumes  144 ,  146 . In embodiments, the ports  147  include a first distal port  148 A for the first actuation volume  144  and a second distal port  148 B for the second actuation volume  146 . The hydraulic drive  110  is operatively connected to one or more of these actuator volumes  144 ,  146  through one or more of the plurality of ports  147 . The hydraulic drive  110  is configured to supply work oil or vent work oil as required by the operating conditions of the compressor module  100 . In some embodiments, one or more main stage valves  250  (“MSV  250 ”) control the flow of work oil to or from one or more of these ports  147  and thereby control the flow of work oil to or from a respective actuation volume  144 ,  146  (see, e.g.,  FIG.  14   ). It will be appreciated that in embodiments any one or more of the of the plurality of ports  147  may be a plurality of ports arranged around the actuator piston  126 , for example the cross-sectional view of  FIG.  7    illustrates two each (at top and bottom) of the first and second distal ports  148 A,  148 B along with two each of the first and second proximal ports  148 C,  148 D. In certain embodiments, the plurality of first and second distal ports  148 A,  148 B are arranged around the actuator piston  126 . The plurality of first and second distal ports  148 A,  148 B may be arranged annularly and/or symmetrically about the actuator piston axis  208 . In embodiments, the drive housing  114  comprises one or more manifold ports  117  for connecting the plurality of ports  147  to exterior components (e.g., a valve manifold  244  discussed below with reference to  FIG.  20   ). 
     As shown in  FIG.  16   , in some embodiments, four MSVs  250 A-D are provided as two for each of the first and second actuation volumes  144 ,  146 , each MSV corresponding to a pressurized circuit of the one or more pressurized circuits  120 . In this embodiment, for the first actuation volume  144 , the MSV  250 C controls a medium-pressure circuit  132  and the MSV  250 A controls a high-pressure circuit  134 ; for the second actuation volume  146 , the MSV  250 D controls the medium-pressure circuit  132  and the MSV  250 B controls the high-pressure circuit  134 . In another sense of this embodiment, each pressure circuit comprises two MSVs  250 , one for supplying work oil and one for venting work oil during a piston stroke, with those roles reversed during the opposite stroke. For example, during the discharge stroke of compressor head  31 , MSV  250 B provides a supply of high pressure work oil through the high-pressure circuit  134 , while MSV  250 C vents work oil on the other side of the actuator piston  126  out of the first actuation volume  144 . 
     As shown in  FIGS.  2 - 8   , in embodiments, the compressor module  100  comprises a first diaphragm compressor head  31  and a second diaphragm compressor head  51  that are each aligned and centered on a compressor axis  206  extending through the center of the diaphragm  5 . In certain embodiments, the first diaphragm compressor head  31  and the second diaphragm compressor head  51  are driven by a single hydraulic actuator  114 . In some embodiments, the hydraulic actuator  114  is operatively coupled to both the first and second diaphragm compressor heads  31 ,  51 , such that the suction cycle of one compressor head aids in initiating the discharge cycle of the other compressor head, which creates a force couple between the compressor heads as discussed further below. 
     In certain embodiments, the process gas discharged from a compressor head (e.g., first compressor head  31 ) is at a relatively low pressure and, for further pressurization, may subsequently be fed into another compressor head, which may be either the second compressor head  51  of the same compressor module  100 , or a compressor head  31 ,  51  of a separate compressor module  100 B-D of the same stack  201 , or a compressor head  31 ,  51  of a compressor module of a separate stack for further compression. 
     In some embodiments, the compressor module  100  is arranged compactly and therefore requires specific hydraulic routing and high pressure gas plumbing and connections. In some embodiments, the compressor heads  31 ,  51  can accommodate reorientation of the inlet and outlet ports  7 ,  9 . As shown in  FIG.  2   , the 180° opposing inlet port  7  and outlet port  9  can be clocked in almost any desired orientation as indicated by the arrows A. 
     In certain embodiments, an energy recovery mechanism can be provided through a force couple architecture, embodiments of which are shown in  FIGS.  2 - 8 ,  9 , and  10 - 11   . Referring to  FIGS.  2 - 8   , some embodiments of this architecture comprise a pair of opposing diaphragm compressor heads  31 ,  51  both driven by an actuator piston  126  that is a double acting double rod, which may or may not act as a hydraulic intensifier, and which is actuated to provide high pressure work oil to actuate the diaphragm compressors. The two pressurized actuation volumes  144 ,  146  are alternately fed pressurized fluid and vented to drive the actuator piston  126  back and forth towards either compressor head  31 ,  51 . Additionally and as discussed further below, since the respective diaphragms  5  of the compressor heads  31 ,  51  oppose each other and are out of phase in this embodiment, the force imposed on one diaphragm by the intake of process gas (e.g., intake of process gas to compressor head  31 ) consequently imposes an aiding force during the opposing diaphragm&#39;s compression and discharge stroke (e.g., compression and discharge from compressor head  51 ). The force couple architecture imposes a force couple to the actuator  114  reducing the force and energy requirements for moving the actuator piston  126  to actuate the diaphragms  5  of both compressor heads  31 ,  51 . 
     For a discharge cycle of the compressor head  31 , operation begins when the actuator piston  126  is at or near the end of its stroke away from the compressor head  31 . At this point, process gas at the inlet pressure has already been supplied to the process gas region  36  of the diaphragm compressor head  31  whereas the opposing second compressor head  51  is fully evacuated of process gas. When diaphragm  5  motion is desired for the compressor head  31 , the MSV  250  actuates to supply pressurized work oil to the second actuation volume  146  on the second side  145  of the actuator piston  126 , forcing the actuator piston  126  up towards the compressor head  31  that is filled with process gas (“up” and other such directions are in reference to  FIG.  6    for sake of clarity and are an example embodiment of the relative movement and positions of various parts, but are not intended to be limiting). As the actuator piston  126  moves, the diaphragm piston  3  pressurizes the work oil in the work oil region  54  below the diaphragm  5 . Since this hydraulic pressure in the work oil region  54  is greater than the pressure of process gas, the diaphragm  5  moves upwards thereby pressurizing the process gas. Once the process gas pressure reaches a target process gas pressure, the process gas is expelled out of the compressor head  31  and either supplied to the tank  256  or supplied to a subsequent compressor head (e.g., the compressor head  51  of the same compressor module  100 , a compressor head  31 ,  51  of another compressor module in the same stack  201 , or a compressor head  31 ,  51  of another compressor module in another stack) for further pressurization. After all or most of the process gas has been forced out of the process gas region  36 , the MSV  250  stops providing hydraulic flow and the actuator piston  126  stops actuating. 
     When diaphragm motion is desired in the opposing direction (i.e., a discharge stroke of the second compressor head  51 ), the MSV  250  is actuated to provide pressure to the opposing first side  143  of the actuator piston  126  into the first actuation volume  144 , thereby forcing the actuator piston in the opposite direction and compressing the gas in the second variable volume region  142  toward the second compressor head  51 . As the hydraulic actuator  112  pressurizes the process gas within the second compressor head  51 , the compressor head  31  is undergoing its intake or suction stroke where the process gas at inlet pressure is supplied above the diaphragm  5  in the process gas region  36 . This initial supply of inlet-pressure process gas may initially assist in providing pressure and moving the diaphragm  5  downwardly and pressurizes the remaining work oil below the diaphragm  5  in the variable volume region  54 , which applies a force to the diaphragm piston  3  thereby providing an aiding force during the opposing compressor head  51  compression, or discharge stroke. This aiding force from the process gas supply reduces the required force from the HPU  118  to drive the actuator piston  140  and compress gas in the second compressor head  51 . Subsequently, process gas completely fills the compressor head  31  in process gas region  36 . To finish the discharge stroke of the second compressor head  51 , pressurized process gas is discharged from the process gas region  36  of the second compressor head  51 . Upon completion of the discharge stroke of the second compressor head  51 , the compressor head  31  is filled with process gas and the second compressor head  51  is fully evacuated of process gas. 
     Referring to  FIG.  9   , another embodiment of a compressor module  100  includes a first and second internal porting  127 A,  127 B through the respective first and second sides  143 ,  145  of the actuator piston  126 . The first side  143  of the actuator piston  126  comprises a first opening  154  and the first internal porting  127 A that are in fluid communication with both the first actuation volume  144  and the first proximal port  148 C. In this manner, the first actuation volume  144  can be supplied or vented through the first internal porting  127 A and the first opening  154 . In the illustrated embodiment, the first proximal port  148 C is part of a low-pressure circuit  130  and is controlled by a main stage valve  250  (“MSV  250 ”) to selectively supply low-pressure work oil to the first actuation volume  144  or vent work oil from the first actuation volume. In other embodiments, the first internal porting  127 A may be in fluid communication with any one or more of the plurality of ports  147  and operable with any one or more of a plurality of pressurized circuits  120 . 
     In embodiments, at least one of the first opening  154  and the first internal porting  127 A of the actuator piston  126  comprises a check valve (not shown) to prevent the flow of work oil out of the first actuation volume  144  through the first internal porting  127 A when the first actuation volume is pressurized for a discharge stroke of the second diaphragm piston  140 , the check valve thereby maintaining the pressure in the first actuation volume  144 . As discussed below, in some embodiments a landing orifice  107  connects the first proximal port  148 C to the first distal port  148 A, and vented work oil from the first internal porting  127 A flows out through the first distal port  148 A via the landing orifice to a pressurized circuit, accumulator, or the reservoir  230 . In some embodiments, additional ports (e.g., first proximal port  148 C in  FIG.  9   ) of the plurality of ports  147  are in fluid communication with the first actuation volume  144  separately from or in addition to the first opening  154 . It will be appreciated that, in embodiments, the first opening  154  and the second internal porting  127 B are provided at the second side  145  of the actuator piston  126  and in communication with the second actuation volume  146  in a substantially similar manner as at the first side  143  of the actuator piston. 
     Referring to  FIGS.  10 - 11   , still another embodiment of a compressor module  100  is shown that is generally similar to  FIG.  9   . In embodiments, the first and second compressor heads  31 ,  51  comprise an oil distribution plate  55  including an array of passages  56  from the respective variable volume region  54 ,  142  to the diaphragm  5 . 
     Referring to  FIG.  10   , in some embodiments, the compressor module  100  comprises a feedback mechanism  108  configured to determine one or more of a position and velocity of the actuator piston  126  during use. The feedback mechanism may include one or more of a sensor  158  and a pressure sensor  159 . In some embodiments, the actuator piston  126  comprises an indication feature  156  that is detectable by the sensor  158 . In various embodiments, the sensor  158  is one or more of an inductive sensor, an optical sensor, a Hall Effect sensor, or the like. 
     In certain embodiments, the indication feature  156  is a variable-geometry portion of the actuator piston  126 , for example a decreasing radius, and the sensor  158  is an inductive proximity sensor configured to measure the distance to the indication feature  158 , the distance measured in a direction perpendicular to the motion of the actuator piston  126  along the actuator piston axis  208 . In one such embodiment shown in  FIG.  10   , as the actuator piston  126  moves right-to-left toward the first compressor head  31 , the sensor  158  can detect an increase in the distance to the indication feature  158  because the radius of the actuator piston is decreasing. Based on the measured distance between the sensor  158  and the indication feature  158 , the feedback mechanism  108  is configured to determine the absolute position of the actuator piston  126 . In embodiments, the feedback mechanism  108  is configured to determine the velocity of the actuator piston  126  based on multiple measurements by the sensor  158  over time. 
     In embodiments, the feedback mechanism  108  comprises a pressure sensor  159  ( FIG.  15   ) operatively coupled to pressurized process gas in or from the compressor head  31 , for example directly measuring pressure of the process gas in the process gas region  36  or measuring the pressure of discharged process gas from the first compressor head  31  after the inlet port  7 . The feedback mechanism  108  is configured to calculate the velocity of the actuator piston  126  based on the measured pressure of the discharged process gas. In embodiments, the feedback mechanism  108  is configured to calculate the velocity of the actuator piston based on multiple inputs, such as measurement(s) from the sensor  158  in conjunction with the pressure sensor  159  or other sensors operatively configured to sense or detect a portion of the compressor module  100  and/or hydraulic drive  110  (e.g., pressure sensor(s) in the first and/or second variable volume region  54 ,  142 , pressure sensor(s) in the first and/or second actuation volume  144 ,  146 ). The feedback mechanism  108  is configured to control other aspects of the module  100  based on the position and/or velocity of the actuator piston  126 , for example controlling the main stage valves  250  to supply or vent work oil to the hydraulic drive  110  or controlling the supply of process gas to a compressor head  31 ,  51 . 
     Referring to  FIG.  19   , an alternative embodiment of a compressor module  100  is shown with the hydraulic drive  110  positioned offset from one or more compressor heads  31 ,  51  with only a hydraulic passage manifold  114 B between the compressor heads. The hydraulic passage manifold  114 B provides passages that hydraulically connects the hydraulic drive  110  to the compressor heads  31 ,  51  without pistons and is significantly smaller than the drive housing  114  of other embodiments. This arrangement reduces the axial length of the stack  201  and may reduce the overall footprint of the stack  201  or the entire high-throughput compressor system  200 . In the illustrated embodiment, the compressor head axis  206  is perpendicular to the actuator piston axis  208 . However, compressor head axis  206  can be oriented in nearly any relationship to the actuator piston axis  208  so long as they are in fluid communication. 
     In embodiments, the compressor heads  31 ,  51  of the compressor module  100  may be independently operated and timed to be synchronized, not synchronized, or alternating. Such arrangements are generally achievable in any compressor architecture that is not force coupled. In embodiments, the compressor heads  31 ,  51  are discharged at substantially the same time. Similarly in embodiments of a stack  201  of compressor modules  100  or a stage  202  of compressor modules  100 , the timing of discharge cycles for compressor heads  31 ,  51  may be independent or dependent within each module, stack, or stage. In embodiments providing independent operation of the compressor heads  31 ,  51 , one or more actuator pistons  126  are separately provided for each compressor head. In certain such embodiments, one or more ports of the plurality of ports  147  are dedicated to a given individual compressor head  31 ,  51  for control of the respective compressor head. In any of the above embodiments with independent operation, any one or more compressor head  31 ,  51 , compressor module  100 , or stack  201  may be selectively turned off and on during operation of the diaphragm compressor system  200 , for example turned off when not needed during certain stages of filling the tank  256 . 
     Pressure Rails 
     The hydraulic system pressure(s) provided by the hydraulic power unit  118  (“HPU”) in some embodiments ranges from 0-5000 psi, but in other embodiments a higher hydraulic pressure is implemented. The HPU  118  in embodiments comprises a single pump/motor, many small pump/motor systems, or fewer larger pump/motor systems, or combinations thereof, as based on operational requirements. In embodiments, the hydraulic drive  110  comprises actively-controlled pressure-compensated pumps or the like in order to actively control hydraulic pressure throughout operating modes. This active control enables the hydraulic drive  110  to operate efficiently by minimizing energy expenditure to meet system requirements. The HPU  118  is configured to provide work oil at a pressure to the drive cavity  116 , and in some embodiments, this pressure is intensified, e.g., by increasing the supply area relative to the piston area. 
     For some embodiments, in order to minimize hydraulic energy consumption, a variable pressure architecture of the compressor module  100  provides a variable-pressure supply of work oil to provide step or analog changes in the applied pressure to any actuator piston  126  as discussed in U.S. patent application Ser. No. 17/522,896. Accordingly, in embodiments, for different operating modes, the hydraulic drive  110  may supply work oil at multiple different set pressures (also referred to as pressure circuits  120  or pressure rails) and/or flowrates. The plurality of pressure circuits  120  comprises one or more low-pressure circuits  130 , medium-pressure circuits  132 , and high-pressure circuits  134 . The term “circuit” is intended to broadly include both the pressurized fluid and the associated structures conveying and controlling the fluid, and one skilled in the art will appreciate that the same structure (e.g., plumbing) may serve as a part of multiple circuits depending on the operating conditions. 
     In some embodiments with multiple pressure circuits, the HPU  118  uses discrete pump/motor sets producing discrete pressures that supply some or all of the plurality of pressure circuits  120  individually in order to eliminate throttling losses. In embodiments, the HPU  118  comprises a variable pump-motor set that is configured to change the speed or pressure of pressurized work oil output by the HPU. In embodiments, the HPU  118  is automatically variable and/or actively controlled, for example, controlled and adjusted in response to conditions in the hydraulic drive  110 , conditions of the outlet process gas or conditions in the fill tank  256 . Moreover, in certain embodiments, any of the above approaches is used to charge one or more accumulators  136  that are included in one or more of the plurality of pressure circuits  120 . 
     In embodiments of the variable pressure architecture, a low-pressure circuit  130  is implemented to provide a “backfill” or “assist” hydraulic supply to the hydraulic system  100  when a higher pressure is not needed (e.g., when ambient-pressure work oil or other relatively low-pressure work oil is sufficient). In certain embodiments, as the hydraulic actuator starts to move from the end of its stroke, the force imposed by the intake stroke process gas on the diaphragm  5  imposes an aiding force on the diaphragm piston  3  and consequently on the actuator  112 . In some embodiments, particularly when a substantially balanced pressure of work oil is applied to the actuator piston  126  (i.e., substantially equal pressures of work oil in the first and second actuation volumes  144 ,  146 , on the respective first and second sides  143 ,  145  of the actuator piston) this force may be enough to move the actuator  112 , or initiate movement of the actuator  112 , with minimal pressure from the HPU  118  or without the addition of hydraulic pressure to available work oil. Such aiding/assist or initiation force may be referred to as “force coupling” of hydraulic drive  110  with the compressor head  31 ,  51 . The drive cavity  116 , however, will still need a supply of work oil to backfill in one of the actuation volumes  144 ,  146  to allow the actuator  112  to move in the opposite direction, which may be provided by the low-pressure supply rail  130 . In embodiments, the low-pressure circuit  130  comprises relatively low-pressure work oil from one or more of the following: unpressurized work oil from the HPU  118 , an oil reservoir  38  of an active oil injection system  30  providing a circuit of supplemental work oil to the compressor heads, vented work oil from the drive cavity  116  in a previous cycle (e.g., intensified work oil vented via a valve and stored in a hydraulic accumulator  136 D as discussed below), vented work oil from the variable volume region  54 , process gas at the inlet pressure, or other sources in the compressor system  100 . 
     In certain embodiments, the one or more pressure circuits  120  comprises a medium-pressure circuit  132  comprising work oil pressurized by the HPU  118  (e.g., by a throttled supply of higher pressure work oil or by a direct supply from one or more pumps/motors of the HPU). In some embodiments, the one or more pressure circuits  120  comprises a high-pressure circuit  134  comprising high-pressure work oil pressurized by the HPU  118 . It will be appreciated that any of the low-pressure circuit  130 , medium-pressure circuit  132 , and high-pressure circuit  134  may be implemented as multiple pressure circuits at different set pressures. The additional circuits of the plurality of pressure circuits  120  allow for finer tuning and control of the compressor module  100 , and increases efficiency by only providing as much pressure as necessary to move the actuator at a particular part of its stroke. 
     As discussed above, in embodiments the compressor module  100  is configured to control the variable-pressure supply of work oil by supplying work oil from the high-pressure circuit  134  after work oil has been supplied from the low-pressure circuit  130  and/or the medium-pressure circuit  132 . In certain embodiments, the hydraulic drive system  110  is configured to control the variable-pressure supply of work oil by sequentially providing work oil to the drive cavity  116  from the low-pressure circuit  130 , the medium-pressure circuit  132 , and the high-pressure circuit  134 . In embodiments with low pressure operating conditions or requirements, it may be sufficient to provide the work oil to the drive cavity  116  from the low-pressure circuit  130  and the medium-pressure circuit  132 , only. 
     In some embodiments, the plurality of pressure circuits  120  are each operatively connected to the drive cavity  116  and may be fed on one or both sides of the actuator piston  126 . In embodiments, the hydraulic drive  110  comprises a passive first valve  131  ( FIG.  17   ) configured to supply work oil from the low-pressure rail of the low-pressure circuit  130  to the drive cavity  116  and an active second valve  133  ( FIG.  16   ) configured to supply work oil from the medium-pressure rail of the medium-pressure circuit  132  to the drive cavity. Certain embodiments further comprise an active third valve  135  configured to supply work oil from the high-pressure rail of the high-pressure circuit  134  to the drive cavity  116 . As detailed below, in embodiments the active second valve  133  and/or the active third valve  135  may be the main stage valve  250  (“MSV”). 
     In certain embodiments, each of the active second valve  133  and the active third valve  135  is configured to adjust from a supply stage to a return stage, the return stage permitting an outflow of intensified work oil from the drive cavity  116  during the discharge cycle of a corresponding one of the compressor heads  31 ,  51 . In embodiments, a hydraulic accumulator  136 D receives the outflow of intensified work oil from the drive cavity  116 . The hydraulic accumulator  136 D in some embodiments operatively functions as a low-pressure accumulator  136 A of the low-pressure circuit  130 , a medium-pressure accumulator  136 B of the medium-pressure circuit  132 , a high-pressure accumulator  136 C of the high-pressure circuit  134 , or a pilot accumulator  136 E of the pilot valve  290 . 
     Supplying flow from the low-pressure circuit  130  into the hydraulic actuator  112  can be achieved several ways. In some embodiments, the fluid can be supplied through a hydraulic valve (in place of the passive first valve  131  in  FIG.  17   ) that opens to allow flow into the actuator  112  then closes when higher pressure fluid is required. In other embodiments, the flow can be supplied through a check valve, such as the passive first valve  131  which opens due to low pressure of work oil in the work oil region  35  during a suction cycle and as the hydraulic actuator  112  starts to move. Since this is a passive valve, it does not need to be actuated when relatively higher pressure fluid is supplied to the drive cavity  116  (e.g., from the medium-pressure circuit  132  or the high-pressure circuit  134 ) will force the valve closed. Alternately, a three-way valve can be used to supply low-pressure or high-pressure fluid to the hydraulic actuator  112  and vent from the hydraulic actuator when desired. The vent can be connected to the low-pressure circuit  130  as outlined above. In this scenario, fluid from the low-pressure circuit  130  can back flow through the passive first valve  131  into the hydraulic actuator  112  as the actuator starts to move. 
     In certain embodiments, a medium-pressure circuit  132  is set to a pressure approximately 50% of the high pressure circuit  134 . In other embodiments, a medium-pressure circuit  132  is set to a pressure approximately 40% to 60% of the high pressure circuit  134 . In some embodiments, the high-pressure circuit  134  is set at a pressure of approximately 5,000 psi, the medium-pressure circuit  132  is set to from 2,500 psi to 3,000 psi, and the low pressure circuit  130  is set to approximately 500 psi. In other embodiments, high-pressure circuit  134  is set to a pressure selected from 3,000 psi, 5,000 psi, and 7,500 psi. In some embodiments, at least one of the high-pressure circuit  134  and medium-pressure circuit  32  are controlled by the HPU  118  to be variable from the maximum pressure for each respective rail. In other embodiments, at least one of the high-pressure circuit  134  and medium-pressure circuits  132  are controlled by the HPU to be variable in a range from 0% to 100% of the maximum pressure for each respective rail. In further embodiments, at least one of the high-pressure circuit  134  and medium-pressure circuits  132  are controlled by the HPU to be variable in a range from 50% to 100% of the maximum pressure for each respective rail. 
     In certain embodiments, the compressor module  100  may include two stages, for example a low pressure stage with compressor head  31  and a high pressure stage with second compressor head  51 , as discussed in U.S. patent application Ser. No. 17/522,896. 
     Force Coupling Arrangements 
     Referring to  FIGS.  34 - 39   , in embodiments, the initial or assisted stroke movement of the actuator piston  126  under force coupling may be achieved by approaches referred to as “high-pressure recovery” or “medium-pressure shuffling” that do not utilize a supply of work oil from the low-pressure circuit  130 . In some low-pressure embodiments of force coupling, both of the actuation volumes  144 ,  146  are opened to vent (e.g. opened to the low-pressure circuit  130 ) to substantially equalize the work oil pressure in the drive cavity  116  on both sides  143 ,  145  of the actuator piston  126  after the piston has completed movement in a given stroke direction and reached its end stop (also referred to as dwell state). At this point, the corresponding compressor head  31  or  51  has completed its discharge cycle. As the subsequent supply cycle begins, process gas at the inlet pressure forces the respective diaphragm  5  of that compressor head toward its first position. When the pressure is substantially equalized in the actuation volumes  144 ,  146 , the force of filling process gas on the diaphragm  5  can take effect to initiate and/or assist movement of the actuator piston  126  in the opposite stroke direction. Then as the actuator piston  126  moves due to this force couple, one volume ( 144  or  146 ) expands and receives work oil from the low-pressure circuit  130  while the other volume compresses and vents work oil to a vent line (e.g., venting to the low-pressure circuit  130  or to the reservoir  230 ). 
     Embodiments of force coupling with low pressure backfill may be at risk of cavitation. Both the high-pressure recovery and medium-pressure shuffling approaches achieve a state of substantially equal work oil pressure on both sides  143 ,  145  of the actuator piston  126 , but do so at greater pressures that reduce or eliminate the risk of cavitation and/or reduce compressibility losses. Additionally, such approaches simplify the design of the hydraulic drive  110 , for example by reducing the number of components such as bypass check valves (not shown) for the low-pressure circuit  130 . 
     It will be appreciated that for any of the embodiments and arrangements described herein, the supply or venting of work oil from the first and second actuation volumes may be achieved by various connections and plumbing components including one or more of the plurality of ports  147  (see, e.g.,  FIGS.  7 - 8   ). Such connections and plumbing components may be operatively connected to one or more of the plurality of pressure circuits  120  via one or more MSVs  250 . In some embodiments and as shown in  FIGS.  34  and  37   , the first actuation volume  144  comprises a first port  148 A and the second actuation volume  146  comprises a second port  148 B. In embodiments, the hydraulic drive  110  comprises dual-pressure plumbing  430  that is configured to connect to two or more pressure circuits of the plurality of pressure circuits  120 . In embodiments, the medium-pressure circuit  132  comprises a MP-rail connector  432  that is arranged to connect the medium-pressure circuit to one or more MSVs  250 . In embodiments, the high-pressure circuit  134  comprises a HP-rail connector  434  that is arranged to connect the high-pressure circuit to one or more MSVs  250 . 
     As shown in  FIGS.  34 - 36   , a “high-pressure recovery” approach to balancing pressures in the drive cavity  116  enables initial movement of the actuator piston  126  under force coupling. In the high-pressure recovery approach, both of the actuation volumes  144 ,  146  are open to a supply of work oil from the high-pressure circuit  134  to substantially equalize the work oil pressure on both sides  143 ,  145  of the actuator piston  126 , allowing the force couple to take effect. Then as the actuator piston  126  moves due to the force couple, one volume ( 144  or  146 ) expands and receives work oil while the other volume compresses and vents work oil. In some embodiments, this movement of work oil does not include additional work oil from the high-pressure circuit  134 . In embodiments, this work oil is vented to the high-pressure circuit  134  and not the low pressure circuit  130 . In other words, the high-pressure recovery approach does not truly “vent” work oil out of the compressing volume, but instead will reuse this work oil in the high pressure circuit  134 , avoid the need to re-pressurize such oil for use from a vent or low pressure circuit  130  back to high pressure circuit  134 . The high-pressure recovery approach may be advantageously implemented for embodiments of the hydraulic drive  110  that do not include any medium-pressure circuit  132 . 
       FIG.  34    shows an embodiment of a high-pressure recovery approach after dwell. Preceding this stage, the actuator piston  126  reaches dwell after compressing to its fullest extent into the second actuation volume  146  and the second actuation volume is open to vent (e.g., in fluid communication with the low-pressure circuit  130 ).  FIG.  36 A  likewise illustrates this dwell state. At the stage of high-pressure recovery in  FIGS.  34  and  36 B , the first actuation volume  144  is closed off from the high-pressure circuit  134 , while both the first and second actuation volumes  144 ,  146  are opened the high-pressure circuit  134 . In  FIG.  36 B , the second actuation volume  146  is also opened to be in fluid communication with the first actuation volume  144 , allowing the volumes to balance the respective pressures. In  FIG.  36 C , the actuator piston  126  moves due to force coupling toward the first actuation volume  144  (i.e., compressing the first actuation volume and expanding the second actuation volume  146 ). During this movement, the second actuation volume  146  expands and receives additional work oil from the shared high-pressure oil of the high-pressure circuit  134 , while first actuation volume  144  compresses and releases high-pressure oil to the high-pressure circuit. In certain embodiments, the force couple only moves the actuator piston  126  a limited distance. In  FIG.  36 D , to continue movement of the actuator piston  126 , the actuation volume  144  is changed to be open to vent (e.g., connected to the low-pressure circuit  130 ) while the high-pressure circuit  134  is closed off, allowing asymmetric chamber pressure to drive the piston. In  FIG.  36 E , the actuation volume  144  is fully compressed and the actuator piston  126  reaches its end-stop (i.e., the next piston dwell).  FIG.  35    shows the valve timing for each of the actuation volumes  144 ,  146  along with the corresponding movement of the actuator piston  126 . 
     As shown in  FIGS.  37 - 39   , initial movement of the actuator piston  126  in certain embodiments is achieved by a “medium-pressure shuffling” approach. In the medium-pressure shuffling approach, similar to the high-pressure recovery, both of the actuation volumes  144 ,  146  of the drive cavity  116  are open to a supply of work oil from the high-pressure circuit  134  to substantially equalize the work oil pressure on both sides  143 ,  145  of the actuator piston  126 . However, the high-pressure recovery approach described above uses the typical plumbing for routing the high-pressure circuit  134  to the drive cavity  116 , which includes an HP-rail connection  434 . By contrast, as shown in in  FIG.  37   , in the medium-pressure shuffling approach, the supply of work oil from the high-pressure circuit  134  is routed differently: through a MP-rail connection  432  (i.e., the plumbing typically used for the medium-pressure circuit  132 ). In embodiments, medium-pressure shuffling at this stage also holds open the medium-pressure circuit  132  while driving the actuator piston  126 . 
       FIG.  37    shows an embodiment of a medium-pressure shuffling approach immediately after piston dwell to begin the next stroke toward actuation volume  144  (i.e., in a first stroke direction). Previous to this point and as shown in  FIG.  39 A , the actuation volume  144  had been open to the high-pressure circuit  134  while the actuation volume  146  was open to vent (e.g., open to the low-pressure circuit  130 ). In  FIG.  21    and  FIG.  39 B , the first and second actuation volumes  144 ,  146  are connected to each other directly and/or via the medium-pressure circuit  132  being routed through the MP-rail connection  432 . In other words, the MSVs  250 B,  250 D are both open to the medium-pressure circuit  132  but are closed off to the high-pressure circuit  134  and the low-pressure circuit  130 . Consequently, both actuation volumes  144 ,  146  substantially balance with work oil being an intermediate-pressure oil  134 ′, i.e., work oil at a pressure between the pressures of the medium-pressure and high-pressure circuits  132 ,  134 . In certain embodiments, the intermediate-pressure oil  134 ′ may be at a pressure approximately the same as the high pressure in the high-pressure circuit  134 , but slightly lower, such as within about 1%, about 5%, or about 10% of the high pressure. In some embodiments, the intermediate-pressure oil  134 ′ quickly equilibrates to a pressure approximately the same as the medium pressure, or slightly above the medium pressure, such as within about 1%, about 5%, or about 10% of the medium pressure in the medium-pressure circuit  132 . In this manner and as shown in  FIGS.  39 B and  39 C , the actuator piston  126  is able to move due to the force couple, the second actuation volume  146  expands and receives additional work oil from the shared intermediate-pressure oil  134 ′, while first actuation volume  144  compresses and releases intermediate-pressure oil to the second actuation volume and/or to the high-pressure circuit  134 . As with other embodiments, the force couple may only move the actuator piston  126  a limited distance. 
     In  FIG.  39 D , to continue movement of the actuator piston  126 , the compressing actuation volume  144  is changed to be open to vent (e.g., connected to the low-pressure circuit  130 ) while the expanding actuation volume  146  continues to be supplied from the medium-pressure circuit  132 , allowing asymmetric chamber pressure to drive the piston. In  FIG.  39 E , the high-pressure circuit  134  further pressurizes the actuation volume  146 . At this stage, the valved connection to the medium-pressure circuit  132  may also stay open due to a check valve  438  ( FIG.  37   ) that prevents high-pressure oil from flowing further up the medium-pressure circuit, which improves subsequent timing when switching back to medium pressure. In  FIG.  39 F , the actuation volume  144  is fully compressed and the actuator piston  126  reaches its end-stop (i.e., the next piston dwell).  FIG.  38    shows the valve timing for each of the actuation volumes  144 ,  146  along with the corresponding movement of the actuator piston  126 . 
     Stack Arrangements 
       FIGS.  30 - 32    show alternative embodiments where some components of the hydraulic drive  110  are shared over multiple compressor heads  31 ,  51  that may be part of a stack  201  applicable to the present disclosure. 
     Referring to  FIG.  30   , in embodiments a common actuator  240  is operatively coupled to several compressor heads, for example compressor heads  31 A-D, while being physically offset from the compressor heads. This arrangement is in contrast to other embodiments with one actuator for each compressor module that is physically housed in the module between compressor heads. The common actuator  240  functions as the intensifier and hydraulic drive for each compressor head  31 A-D. The common actuator  240  in embodiments is driven by a single HPU  118  or multiple HPUs. In certain embodiments and as illustrated in  FIG.  30   , the common actuator  240  provides pressurized fluid to simultaneously actuate the diaphragms  5  of both compressor heads  31 A,  31 B, and then the common actuator  240  reverses directions to simultaneously actuate the diaphragms  5  of both compressor heads  31 C,  31 D. 
     In some embodiments of the stack  201 , the common actuator  240  is mounted in the stack with the actuator piston axis  208  coaxial with the compressor head axis  206 . In other embodiments, the common actuator  240  is separate from the stack  201 . In still other embodiments, the common actuator  240  and the corresponding compressor heads  31 A-D are all separate from any stack  201  to operate as an auxiliary compressor plumbed to another compressor module or stack. In any such embodiments, a single compressor head of the compressor heads  31 A-D may be configured to be taken offline while the remaining compressor heads continue to operate. 
     Referring to  FIG.  31   , in some embodiments a first and second compressor module  100 A,  100 B share a common control valve, MSV  250 . The MSV controls a pressurized supply of work oil from the HPU  118 . In this sense, the HPU  118  and the MSV  250  are configured to supply and control the supply of pressurized work oil to a plurality of compressor modules and a plurality of compressor heads. In other embodiments, multiple MSVs  250  and/or multiple HPUs  118  are provided and shared by the first and second compressor modules  100 A,  100 B, for example when providing both medium-pressure and high-pressure circuits  132 ,  134 . 
     In an embodiment shown in  FIG.  32   , the HPU  118  is configured to act directly on the diaphragm  5  of one or more compressors  31  while omitting hydraulic actuator  112  and the piston subassembly  122 . The MSV  250  is operatively connected to the HPU  118  to control the supply of work oil directly to the diaphragms  5 . In the illustrated embodiment, the MSV  250  controls the supply to three compressors  31 . In embodiments, any one or more of the pressure circuits  120  is implemented and controlled by one or more MSVs  250  for one or more of the compressor heads  31 . Although  FIG.  32    is illustrated schematically, it will be appreciated that the physical arrangement of the compressor heads  31  of the stack  201  can be coaxial on a compressor axis  206  as in previous embodiments. A compressor stack  201  implementing this embodiment provides an axial length and overall footprint are significantly decreased. 
     In any embodiments of the present disclosure, each pressure circuit of the one or more pressure circuits  120  may be independently and actively controlled to adjust the amount of pressure supplied to the hydraulic actuator  112 . In embodiments, the active valves  133 ,  135  or MSV  250  may be controlled to adjust the respective pressure circuit. The plurality of ports  147  may similarly comprise a valve to actively control or throttle the flow to the drive cavity  116 . It will be appreciated that the hydraulic drive  110  is likewise configured for nearly instantaneous stoppage of the actuator piston  126  and shutdown of the compressor module  100  due to the HPU  118  along with the active valves  133 ,  135  and/or the MSV  250 , any associated control mechanisms (e.g., feedback mechanism  108 ), or shutoff valves. For example, the actuator piston  126  can be stopped during a discharge or suction stroke before such stroke is completed by closing off the pressure circuit(s) that are pressurizing the corresponding actuation volume. Accordingly, the hydraulic drive  110  is configured to stop a stroke of the actuator piston  126 , a stroke of the diaphragm piston(s)  3 ,  140 , and/or actuation of the diaphragm(s)  5  before the stroke or actuation completes its current cycle. This shutoff capability provides safety by minimizing further damage when a hazardous condition is detected. This is an improvement over prior compressor drives, e.g., crank-driven systems, which must mechanically stop components and overcome significant inertial forces before stopping, resulting in continuing operation during the hazardous condition. 
     Clamping Mechanism 
     Referring to  FIG.  1   , in embodiments, the high-throughput compressor system  200  comprises a clamping mechanism  204  that holds the compressor modules  100  together while accommodating the significant pressures, vibrations, and other forces experienced during compressor cycles. In other embodiments, a clamping mechanism  304 ,  404 ,  504 ,  604 , or  704  is provided as discussed below, with broad functionality similar to the clamping mechanism  204  (see  FIGS.  24 - 29   ). Generally, each individual compressor head  31 ,  51  is formed of multiple plates that must be clamped together with enough force to resist cyclical forces including pressurized work oil, pressurized process gas, and diaphragm actuation without leakage. As such, a conventional individual compressor head requires a specialized individual mechanism such as a large number of high-strength bolts to sufficiently clamp the head together. By contrast, the clamping mechanism  204  of the present disclosure applies a clamping force sufficient to hold together each such compressor head for multiple compressor modules  100  with minimal or no clamping within individual heads. 
     In certain embodiments, the clamping force is exerted by the clamping mechanism  204  at opposite ends of a stack  201  of one or more compressor modules  100 , with the force acting through each module  100  to clamp all of the heads  31 ,  51  of each module  100  in the stack. Clamping together each head  31 ,  51  comprises clamping together support plates that define each head, resisting pressure of compressed fluid(s) inside the head, and clamping one or more of the support plates (e.g., work oil head support plate  8 ) to a drive housing  114  of the compressor module  100 . It will be appreciated that the clamping mechanism  204  therefore eliminates and/or reduces other hardware, including bolts and also the size and thickness of components of the heads  31 ,  51 , necessary for clamping a conventional compressor head, and may provide reduced assembly time, reduced size and weight for each module  100  compared to a conventional diaphragm compressor, and improved serviceability. The clamping mechanism  204  is therefore also applicable to a single compressor head  31 ,  51  or a single compressor module  100  that is not stacked with other modules. In certain embodiments, the total clamping force necessary to operate each head  31 ,  51  in a stack  201  of modules  100  is not provided by bolts securing each individual head  31 ,  51  to each respective module  100 . In other embodiments, the total clamping force is provided by a combination of the clamping mechanism  204  and bolts securing each individual head  31 ,  51 . 
     In embodiments, the compressor system  200  comprises a clamp actuator  212  that applies a compressive force (i.e., clamping load) to the compressor modules  100  while also accommodating changes in thermal expansion of the compressor modules during operation. If the stack  201  were rigidly clamped without the clamp actuator  212 , significant stresses would arise in the compressor modules  100  due to thermal growth of hardware that results from temperature increases as process gas is compressed. While accommodating thermal expansion, the compressive force of the clamp actuator  212  may be constant or substantially constant. In embodiments, the clamp actuator  212  is a hydraulic load actuator. The present disclosure provides several embodiments that are based around an actuator such as the clamp actuator  212  and a reactionary structure such as a frame or tie rod arrangement. 
     In the embodiment of  FIG.  1   , the clamping mechanism  204  comprises a base plate  220  and an end plate  222  connected by four tie rods  224  with respective tensioner nuts  226 . In some embodiments, the stack  201  is mounted on a skid  228  or similar base. The compressor modules  100  are aligned along a common axis which, as discussed above, is a compressor head axis  206  of each compressor module  100 . In some embodiments, one or both of the base plate  220  and end plate  224  are movable or repositionable along the compressor axis  206  to accommodate thermal expansion or various sizes of compressor modules  100 . In certain embodiments, one or both of the base plate  220  and the end plate  224  are movable by the clamping actuator  212 . 
     Referring also to  FIG.  21   , the clamp actuator  212  in embodiments is a hydraulically-powered piston actuator comprising a hydraulic cylinder  232  and a piston  233 . In the illustrated embodiment, the hydraulic cylinder  232  would have pressure applied to create the required clamping force for the constituent compressor modules  100  and compressor heads  31 ,  51 . The same clamp actuator  212  architecture could be used for all compressor head sizes, but with different clamping loads as required for different supply pressures. In embodiments, the hydraulic supply for the clamp actuator  212  may be shared with one or more of the compressor modules  100 , for example from a hydraulic power unit  118  of a compressor module. 
     In embodiments, operation of the clamp actuator  212  comprises manually charging the hydraulic cylinder  232  and subsequently monitoring the pressure within the clamp actuator  212 . The cylinder  232  is then resupplied if pressure drops and leakage occurs over time, for example through dynamic seals. Alternatively, the hydraulic circuit of the clamp actuator  212  could be supplied with an additional make-up pump (not shown) to accommodate the lost fluid. The make-up pump may be similar to, or the same as used for the active oil injection system  30  applicable to embodiments of the compressor module  100  discussed below, and would be an adequate option to provide a low flow high pressure supply source. In embodiments, the monitoring may be automated with a configuration where the compressor system  200  shuts down when the pressure within the clamp actuator  212  drops below a certain threshold, and oil may also be injected, or oil pressure increased, as detected by the system. 
     In certain embodiments, as one mode to accommodate thermal growth, the large volume of oil within the hydraulic cylinder  232  of the clamp actuator  212  inherently has some compressibility, which may account for some of the increased pressure from modules  100  due to thermal growth. Even with this compressibility, the clamp load could grow by nearly 25% with a four-module stack  201 . This load increase is proportional to the length of the overall stack  201 . To further reduce the stiffness of the volume, some embodiments incorporate a piston accumulator (not shown), for example a high pressure piston accumulator that can accommodate pressures up to 1,000 bar. Moreover, such an accumulator also allows additional time for system shut down if a seal were to fail. In some embodiments, the pressure is monitored, and a lower threshold is set such that when the pressure drops below the set point, the system shuts down and sends an alarm to the operator. The lower set point, however, is still within a reasonable clamp load such that the compressor modules remain under load during the shutdown. Since the compressor modules are hydraulically driven, the shutdown may be rapid. 
     For service and change out of compressor modules  100 , once any plumbing or electrical connections are disconnected from a given compressor module  100 , the individual compressor module can be deactivated, serviced, and/or removed as a unit. In embodiments, an overhead gantry crane (not shown) is integrated into the stack skid  228 , and has a designated area at the end or off to the side. In embodiments, each compressor module  100  comprises an eye bolt  246  ( FIG.  2   ) or other attachment for lifting and moving the module via the gantry crane or the like. In some embodiments, each compressor module  100  comprises one or more feet  242 , such as four feet shown in  FIG.  3   , which may function as a cart to move the module and/or may function to engage the skid  228  or other base of the stack  201 . 
     Referring to  FIG.  24   , in other embodiments, a clamping mechanism  304  comprises a base plate  320  and end plate  322  that are connected by two tie rods  324 . The clamp actuator  212  may similarly be provided between the base plate  320  and the compressor modules  100 . Both of these embodiments of the tie rod based clamping mechanisms  204 ,  304  have their own respective advantages. The four-bolt arrangement of clamping mechanism  204  can allow for easier service (e.g., service in place, change out, or temporary removal for service) of the compressor modules  100  but may be more difficult to accommodate plumbing. The two-bolt arrangement of clamping mechanism  304  provides easier plumbing access but may be more difficult to service a module. 
     Referring to  FIG.  25   , in still other embodiments, a clamping mechanism  404  comprises a base plate  420 , an end plate  422 , and a reactionary frame  425  mounting the plates along with the compressor stack  201  and a clamp actuator  412 . In some embodiments, the actuator  412  may be the same as the clamp actuator  212 . In embodiments, the reactionary frame  425  is rigidly affixed (for example, bolted) to a foundation. In some embodiments, to withstand the large tensile loads applied, the reactionary frame  425  is a one-piece unitary component or two pieces rigidly affixed together, and in particular embodiments the reactionary frame is formed of one cast metal part or two cast metal parts rigidly affixed together. As shown in  FIG.  26   , this embodiment of clamping mechanism  404  leaves the top and sides of the compressor modules  100  substantially open and accessible, which provides access for plumbing and servicing. However, the overall size of the clamping mechanism  404  may be larger than other embodiments. 
     Referring to  FIGS.  26 - 27   , a clamping mechanism  504  provides a different embodiment based on the four tie rod embodiment. The clamping mechanism  504  comprises a base plate  520  and end plate  522  that are connected by four tie rods  524 , and a clamp actuator  512  mounted to the base plate. In this embodiment, pre-tensioning nuts  526  are mounted on the tie rods  524  at the base plate  520  and apply an initial preload to the stack  201  before operation. The remainder of the required clamping force is then made up by the clamp actuator  512 . As the clamping actuator  512  pressure is applied, a thermal expansion gap  513  is created at the base of the actuator to accommodate thermal expansion. 
     One benefit of this embodiment is that the pre-tensioned of the tie rods  524  creates a safety if the clamp actuator  512  fails such that the stack  201  is contained to the initial preload. However, service of each compressor module  100  may be more challenging than some of the other embodiments since the tie rods  524  are mounted more closely to the compressor modules. 
     Referring to  FIG.  28   , an embodiment of the clamping mechanism  604  provides a different embodiment based on the four tie rod embodiment. The clamping mechanism  604  comprises a base plate  620  and end plate  622  that are connected by four tie rod assemblies, each including a first tie rod  624 A and a second tie rod  624 B connected by a coupler  623 . In embodiments, an additional plate  621  is mounted inside the base plate  620 . In this embodiment, the coupler  623  provides for ease of assembly and service by separately receiving the first and second tie rods  624 A,  624 B, which allows for the first and second tie rods to be relatively shorter and able to be installed from each respective side of the stack instead of assembling one long tie rod across the entire stack. In embodiments, the coupler  623  is a threaded collar. In other embodiments, the coupler  623  may provide some or all of the clamping load and thermal accommodation. . In the illustrated embodiment, the tie rods  624 A-B are entirely outside of the modules  100 , providing easier access for lifting the modules out of the stack. 
     Referring to  FIG.  29   , an embodiment of the clamping mechanism  704  provides a variation on the four tie rod concept. The clamping mechanism  704  comprises a base plate  720  and end plate  722  that are connected by four tie rods  724 . In this embodiment, Belleville washers  730  are mounted on the tie rods  724  at the end plate  722  and to apply a clamping load to the stack  201 . The Belleville washers are springs that are mounted to be biased in the direction of the clamping force. In use, the Belleville washers  730  receive the axial load from thermal expansion and may compress under this load while maintaining the requisite clamping force against the stack  201 . For stacks  201  of a different number of size of compressor modules  100 , the clamping mechanism  704  may be adjusted by selecting Belleville washers  730  of different sizes or a different number stacked. Additional clamping force is provided by the clamp actuator  712 . 
     Another embodiment of the present disclosure incorporates a hydraulic actuator (such as clamp actuator  212 ) to apply initial load and a threaded lock ring (not illustrated) to mechanically maintain the load. This eliminates piston seal failure as a failure mode. However, this approach provides limited thermal accommodation, or requires an adjustable lock ring. 
     In other embodiments, the clamp actuator  212  is modified to minimize axial length, reducing the overall footprint of the stack  201 . If the parameters of the high throughput compressor system  200  and the compressor modules  100  are known, the length of the piston  233  and the cylinder  232  can be decreased to a minimum size that is capable of providing the corresponding required reactive clamping force and thermal accommodation. Additional concepts may be employed to reduce overall piston size, such as introducing a lever (e.g., a single lever arm or compound lever) functionally between the piston and the stack  201 , for example with one end of the lever rigidly affixed to the piston and the other end of the lever rigidly affixed to the second compressor head  51  of the first compressor module  100 A. The fulcrum of the lever may be mounted to the skid  228  or incorporated with other fixed parts of the clamping mechanism  204 / 304 / 404 / 504 / 604 / 704 . Such a lever arm multiplies the linear force of a downsized piston. 
     Staging and Reconfigurability 
     Embodiments of the high throughput compressor system  200  provide several potential arrangements of interconnecting and controlling the plurality of compressor modules  100  to customize a tank-filling operation. In various embodiments, some or all of the compressor modules  100  may be operatively arranged in parallel to pressurize a larger volume of process gas than would be accomplished by a single compressor module  100  over a period of time. In some embodiments, some or all of the compressor modules  100  may be operatively arranged in series to progressively pressurize process gas to a higher pressure than would be accomplished by a single compressor module  100  or a single constituent compressor head. In certain embodiments, the high throughput compressor system  200  is controlled and configurable to switch between such modes, or to combine such modes, e.g. some modules  100  operating in series and some modules  100  operating in parallel. For example, many or all of the diaphragms  5  can be used in parallel to start filling a tank  256  at low pressure. In this embodiment, some compressor heads  31 ,  51  may be low pressure heads  31 , and some may be high pressure heads  51 . In this mode, both high  51  and low  31  pressure heads are used to pressurize process gas to the pressure of the low pressure head  31  or less. Then as the pressure rises the system can be reconfigured to be a two-stage compressor (i.e., first and second stages of different pressure in series) by switching valves. In this embodiment, the low pressure head  31  pressurizes gas from a low pressure to a medium pressure, and this medium pressure gas is fed to the high pressure head  51  that pressurizes the process gas from a medium pressure to a high pressure. The heads  31 ,  51  can be configured such that the compressor system  200  can have as many stages as are desired for optimizing the tank fill process. 
     In embodiments, the compressor modules  100  are mechanically driven, e.g. with a cam driven shaft. In other embodiments, the compressor modules  100  are hydraulically driven. In order to provide high throughput and high pressure of process gas under the physical constraints of hydraulic power and diaphragm actuation, the compressor modules  100  of the present disclosure are configured to operate at high speeds with precision to prevent damage to any components. To this end, embodiments of the compressor modules  100  implement fast valving, particularly in a main stage valve  250  (“MSV  250 ”) controlling the hydraulic supply, and damping of moving components such as pistons or valves. 
     In embodiments, one or more compressor modules  100  may have compressor heads  31 ,  51  that are different from the heads of other compressor module(s) with regard to pressure discharge, throughput, and/or size. In some non-illustrated embodiments, a compressor module  100  may have compressor heads  31 ,  51  that are different from each other, or a different number of compressor heads such as one, three, four, or more compressor heads. 
     Referring to  FIG.  22   , an embodiment of the high-throughput compressor system  200  includes two stages  202 A and  202 B. In certain embodiments, components of the compressor system  200  are configured for process gas inlet at 120 bar. In some embodiments, each stage  202 A-B is split into two stacks  201 A-B and  201  C-D respectively; each stack  201 A-D comprising four compressor modules  100 A-D, eight compressor heads  31 ,  51 , and four hydraulic drives  110 . In certain embodiments, each hydraulic drive  110  has a dedicated HPU  118  with a single pump/motor combination or a dedicated pump/motor group. This arrangement is advantageous from an operational flexibility perspective, in that an individual hydraulic drive  110  (serving two compressor heads) can be taken offline, and its corresponding HPU  118  pump(s)/motor(s) turned off completely, while the rest of the compressor system  200  continues to operate. In one embodiment of a two-stage, single pressure supply scenario, each HPU  118  comprises a 250 hp motor and a pump sized at either 270 or 360 cc/rev. The flow requirements may be satisfied with a 2″ ID supply rail hose, and return rails can be joined into a single large, welded pipe. 
     Referring to  FIG.  23   , another embodiment of the high-throughput compressor system  200  is illustrated with four stages  202 A-D of increasing pressure output and stage bypassing to allow reconfigurability. In some embodiments, three bypass valves  257  are implemented respectively preceding the second, third, and fourth stages  202 B-D to selectively bypass these latter stages. With stage bypassing, when the tank pressure is low, only the lower pressure stages may be doing the compression work, and in one embodiment the discharge process gas may bypass the upper stages. This embodiment avoids pressure drop through the additional lines, check valves and intercoolers of the upper stages, improving efficiency. In embodiments, the bypass valves  257  are three-way valves or an equivalent combination of 2-way valves. It will be appreciated that the compressor system  200  may operate when the tank  256  pressure is at any level within the operating parameters, i.e., from 0-100% of the target tank fill pressure. Accordingly, the compressor system may begin filling a higher pressure tank  256  with the high pressure stages. 
     In the illustrated embodiment of  FIG.  23   , the first stage  202 A is the lowest pressure stage and comprises six compressor heads  31 ,  51  arranged as three compressor modules  100 A-C, with each compressor head configured to pressurize process gas to 1,000 psi. The second stage  202 B is the second lowest pressure stage and comprises three compressor modules  100 A-C totaling six compressor heads  31 ,  51 , with each compressor head configured to pressurize process gas to 2,000 psi. The first and second stages  202 A,  202 B may be in separate respective first and second stacks  201 A,  201 B or arranged in a single stick  201 A-B. The third stage  202 C is the second highest pressure stage and comprises sixteen compressor heads  31 ,  51  arranged as eight compressor modules  100 A-H in one or two stacks  201 C, with each compressor head configured to pressurize process gas to 7,500 psi. The fourth stage  202 D is the highest pressure stage and comprises eight compressor modules  100 A-H in one or two stacks  201 D totaling sixteen compressor heads  31 ,  51 , with each compressor head configured to pressurize process gas to 15,000 psi. Generally, the stages  202 A-D may be referred to by their relative output pressure. One or both of the first and second stages  202 A-B may be considered “low pressure stages” while one or both of the third and fourth stages  202 C-D may be considered “high pressure stages;” alternatively, the second and third stages  202 B-C may be considered “medium pressure stages.” In some embodiments, the high-throughput compressor system  200  may increase the overall process gas throughput by including plumbing for a given compressor head  31 ,  51  to be used for multiple stages, providing selective gas configurability. Although the lower pressure heads  31 ,  51  may not be capable of use as high pressure stages when the gas discharge pressure is high, the higher pressure heads  31 ,  51  and components (e.g., high-pressure circuit  134  and/or medium-pressure circuit  132 ) can be used as a lower pressure stages when the process gas discharge pressure of the system is low, and the overall pressure increase does not yet require as many compression stages. The reconfiguration of the gas compression heads to serve as different stages may provide an increase in flow throughput and consequently e.g. reduce tank filling times. 
     In some embodiments such as  FIG.  23   , the high-throughput compressor system  200  comprises four check valves  258  to implement the gas configurability option. This configuration allows the second stage  202 B to also selectively function as the first stage (e.g., outputting process gas at 1,000 psi), the third stage  202 C to also selectively function as the second stage (e.g., outputting process gas at 2,000 psi), and the fourth stage  202 D to also selectively function as the third stage (e.g., outputting process gas at 7,500 psi). This embodiment may result in a gain of up to about 15% in flow throughput over the course of a tank fill compared to sequentially running the stages  202 A-D individually (e.g. 11.5 kg/min vs 10 kg/min). When the outlet pressure is low (e.g. 30 to 85 bar for hydrogen process gas), only a single compression stage is needed, so all of the compressor heads  31 ,  51  may be plumbed in parallel and configured to provide the pressure of the first stage  202 A. When the outlet pressure increases (e.g. 85 to 250 bar), only two compression stages are needed, and roughly ¾ of the total compressor displacement may be used as the first stage  202 A, and the remaining ¼ may be used as the second stage  202 B. The increased flow rates of this embodiment are most significant when the system can operate as 1 or 2 stages, but this may only possible for about 5% and 20% (respectively) of the total duration of the tank fill. Generally, the compressor heads  31 ,  51  of any stage  202 A-D are able to provide the selective operation at other desired pressures as long as sufficient pressure is applied to the diaphragm  5 ; in certain embodiments this pressure is due to the suction condition applied to the compressor head, for example the suction pressure defined by the check valve  258  at the outlet port  7  (see also  FIGS.  6 ,  15   ) 
     It will be appreciated that embodiments of the present disclosure may comprise various numbers and physical arrangements of stages  202 , stacks  201  per stage  202 , compressor modules  100  per stack  201 , compressor modules  100  per stage  202 , compressor heads  31 ,  51  per compressor module  100 , and compressor heads  31 ,  51  per stage  202 . Moreover, embodiments may have various pressure ratings of the compressor heads  31 ,  51 . Accordingly, embodiments of the diaphragm compressor system  200  comprise from one to ten stages  202  with particular embodiments comprising one, two, three, four, five, six, seven, eight, nine, ten or more stages  202 . Embodiments of the diaphragm compressor system  200  comprise from one to ten stacks  201  or more and any number of the stacks may comprise one or more stages  202  (for example, stack  201  in  FIG.  1    may be configured with compressor modules  100 A-B comprising a first stage  202 A and compressor modules  100 C-D comprising a second stage  202 B). In some embodiments, an individual stack  201  comprises one, two, three, or more stages  202 . Embodiments of the diaphragm compressor system  200  comprise from one to twelve or more compressor modules  100  per stack  201  and, similarly, one to twenty-four compressor heads  31 ,  51  per stack, or any ranges therebetween. Other embodiments comprise stacks  201  with different numbers of compressors  100  or compressor heads  31 ,  51  per stack. Certain embodiments of the diaphragm compressor system  200  comprise from one to six compressor heads  31 ,  51  per compressor module  100 . 
     The output performance of embodiments of the diaphragm compressor system  200  may likewise have various configurations system-wide and various output configurations among the constituent compressor heads  31 ,  51 , modules  100 , and stages  202 . For compressed hydrogen process gas, embodiments of the diaphragm compressor are configured to output pressures up to 30,000 psi or more. Embodiments of the diaphragm compressor system  200  comprise stages  202  that each have a compression ratio in a range of about 1:1 to 10:1 or a range of about 2:1 to 6:1; such ratios may be distinct from each other. In certain embodiments, the compressor system  200  comprises a first stage  201 A outputting process gas at about 40-7,500 psi and an additional stage (e.g., second stage  201 B, third stage  201 C, and/or fourth stage  201 D) outputting process gas at about 1,000-15,000 psi. In other embodiments, the compressor system  200  comprises a first stage  201 A outputting process gas at about 100-7,500 psi, optionally a second stage  201 B outputting process gas at about 200-15,000 psi, optionally a third stage  201 C outputting process gas at about 300-25,000 psi, and optionally a fourth stage  201 D outputting process gas at about 400-30,000 psi. 
     In embodiments, the high-throughput compressor system  200  is configured for a tank-filling operation from 30 to 1000 bar, with a 4-stage system comprising stages  202 A-D. This is generally similar to  FIG.  23   , comprising a first stack  201 A of compressor modules  100  comprising a first stage  202 A of the lowest pressure, a second stack  201 B of compressor modules  100  comprising a second stage  202 B of a higher pressure than the first stack, a third stack  201 C of compressor modules  100  comprising a third stage  202 C of a higher pressure than the second stack, and a fourth stack  201 D of compressor modules  100  comprising a fourth stage  202 D of a highest pressure. 
     Disclosed embodiments and features of the high throughput compressor system  200  for hydrogen process gas can meet a throughput target of up to 10 kg/min or more compressed hydrogen at a minimum outlet pressure of 875 bar, with embodiments capable of 1,000 bar or more. Embodiments provide a compact compressor module  100  that can be stacked together with one or more additional compressor modules and plumbed to achieve essentially any required throughput. For some embodiments, the design is a system with two stages  202 A-B with approximately sixteen diaphragms  5  (each compressor head  31 ,  51  having one diaphragm  5 ) per stage  202 , which may be eight compressor modules  100  per stage and is designed for 120 bar inlet pressure. The present disclosure can be applied to a larger system with lower inlet pressure such as 30-50 bar inlet pressures, requiring four stages  202 A-D to achieve compression up to 1,000 bar. 
     Certain embodiments provide two stacks  201  of four compressor modules  100  per each stage  202 , although other arrangements are feasible and contemplated. Each module  100  may actuate two compressor heads  31 ,  51  of the same size and operating pressures, as would all modules  100  within the same stack  202 , and thus have common suction and discharge gas pressures. This has benefits for simplicity of gas plumbing (e.g., hydraulic components such as HPU  118 , pressure circuits  130 ,  132 ,  134 ,  138  or main stage valve  250  are the same and may be operatively connected to multiple compressor modules  100 ), and also reduction in accumulator count and size (e.g., one or more of accumulators  136 A-E (see, e.g.,  FIGS.  2 - 3 ,  15   ). 
     In some embodiments, one of the compressor modules  100  can be deactivated within the stack  201  and still allow the stack to operate. The compressor module may be deactivated for servicing, repair, or replacement, e.g. by isolating the module  100  by valving. Because the rest of the modules  100  in the stack  201  continue to operate, maintenance can be temporarily deferred for a more convenient time if desired. Additionally, for some embodiments such as embodiments with multiple stacks  201  per stage  202 , the compressor system  200  may provide continued operation during service, albeit at a reduced throughput. This could be achieved by valving and deactivating one stack  201 A while the other stack  201 B of the stage  202 A continues. Consequently, the other stacks  201 A of other stages  202 B before or after the compressor module being serviced may need to be shut down accordingly to match pressure ratios per stage, but this may nonetheless allow continued operation during service and make emergency service requirements less detrimental to overall system operation. Effectively, by having multiple stacks  201  per stage  202 , the system may create a redundancy effect which is beneficial from a failure and service perspective. 
     Damping of the Hydraulic Actuator 
     As shown generally in  FIGS.  6 - 11    with certain embodiments detailed in  FIGS.  8  and  11   , certain embodiments of the compressor module  100  comprises a damping mechanism  105  that includes venting of the work oil being compressed in the direction of travel of the actuator piston  126 . Generally, in some embodiments, the actuator piston travel distance in a discharge stroke may be about 0.5-3 inches, about 1-2.5 in., about 0.5 in., about 1 in., about 1.5 in., about 2 in., about 2.5 in., about 3 inches, or about 0.5 to 4 inches. In embodiments, the travel time of the actuator piston is less than 100 milliseconds (ms). In certain embodiments, the travel time is about 30-95 ms, about 45-75 ms, about 50-70 ms, or about 60-65 ms. Additionally, in embodiments the dwell time of the actuator piston  126  is less than about 50 ms, less than about 25 ms, about 5-30 ms, about 10-25 ms, or about 15-20 ms. Accordingly, the actuator piston  126  reciprocates with quick starts and quick stops including possible impact with a hard stop  106  formed in the drive housing  114 . Embodiments of the present disclosure comprise a damping mechanism  105  to aid in stopping the actuator piston  126  as it approaches the end of a stroke to decrease the impact velocity against the hard stop  106 . The damping mechanism  105  is provided on both sides of the actuator piston  126  and in each of the first and second actuation volume  144 ,  146 . 
     The drive housing  114  comprises a plurality of ports  147  including the first and second distal ports  148 A,  148 B that are in fluid communication with components of the hydraulic drive  110 . The hydraulic drive  110  and the HPU  118  are configured to provide a variable-pressure supply of work oil to the drive cavity  116  through one or more of the plurality of ports  147 . One or more of the plurality of ports  147  is configured to supply work oil to the hydraulic drive  110 , and one or more of the plurality of ports is configured to vent work oil out of the hydraulic drive. In the illustrated embodiments, the plurality of ports  147  is configured to both supply and vent work oil from the hydraulic drive  110  depending on the direction of travel of the actuator piston  126 , as with the first and second distal ports  148 A,  148 B of  FIGS.  6 - 8   . The first and second distal ports  148 A,  148 B are each operatively coupled to a respective main stage valve  250  to control the supply and vent operations. 
     Referring to  FIGS.  6 - 8   , the drive housing  114  comprises orifices  152  for both the first and second actuation volumes  144 ,  146 . The orifices  152  are operatively connected to either a first radial port  153  at the first actuation volume  144  or a second radial port  155  at the second actuation volume  146 . In other embodiments and as shown in  FIGS.  10 - 11   , the damping mechanism  105  comprises a first plurality of orifices  152 A and a second plurality of orifices  152 B in communication with the drive cavity  116 . In embodiments, the orifices  152 ,  152 A,  152 B are also in communication with one or more ports of the plurality of ports  147 . In embodiments, one or more of the orifices  152 ,  152 A,  152 B are in communication with the plurality of ports  147  for both venting and supply of work oil. In certain embodiments, the orifices  152 ,  152 A,  152 B at the first actuation volume  144  are in fluid communication with the first distal port  148 A and the orifices  152 ,  152 A,  152 B at the second actuation volume  146  are in fluid communication with the second distal port  148 B. For any such embodiments, the work oil that is vented from the drive cavity  116  through the orifices  152 ,  152 A,  152 B and one or more of the plurality of ports  147  is supplied to the reservoir  230  or an accumulator such as the recovered oil accumulator  136 D. 
     During the discharge cycle of the first compressor head  31 , the hydraulic drive  110  is configured to provide the variable-pressure supply of work oil through the second distal port  148 B to the second actuation volume  146  to press against the second side  145  of the actuator piston to drive the actuator piston, driving the first diaphragm piston  3  toward the corresponding first compressor head  31 , intensifying the work oil in the first variable volume region  54  to an intensified pressure, and actuating the diaphragm  5  of the first compressor head to the second position. As the actuator piston  126  moves, the damping mechanism  105  comprises the drive cavity  116  being configured to dampen the drive motion of the actuator piston  126  due to a volume of work oil in the opposing first actuation volume  144  with outflow restricted (i.e., the first actuation volume is in the direction of travel of the actuator piston). In certain embodiments, the volume of work oil vents through the first plurality of orifices  152 ,  152 A and out of the first actuation volume  144 , providing space for the actuator piston  126 . Therefore, the damping force of the damping mechanism  105  is a function of the number and size of the plurality of orifices  152 ,  152 A, (and  152 B discussed below), provided that the orifices freely flow to vent. 
     At the beginning of the actuator piston  126  stroke for the discharge cycle of the first compressor head  31 , the first plurality of orifices  152 ,  152 A is open to the first actuation volume  144 . Subsequently, the first plurality of orifices  152 ,  152 A is progressively covered by the actuator piston  126  as it moves along its driving stroke, which constricts outflow through the first plurality of orifices and increases the damping force of work oil remaining in the first actuation volume  144  against the first side  143  of the actuator piston. In other words, as the obstruction by the actuator piston  126  occurs the effective size of the plurality of orifices  152 ,  152 A decreases and the damping force increases because there is less area for the work oil in the first actuation volume  144  to escape. In this manner, the damping mechanism  105  provides an increasing damping force configuration due to work oil that remains in the first actuation volume  144  having access to less available venting area. 
     In some embodiments and as shown in  FIGS.  10 - 11   , the drive housing  114  further comprises a second layer of orifices illustrated as a plurality of second or supplemental orifices  152 B in communication with the first actuation volume  144 , the plurality of supplemental orifices being staggered axially relative to the plurality of first orifices  152 A. In the illustrated embodiments, the plurality of supplemental orifices  152 B are located relatively closer to the respective compressor head  31 ,  51  and further along the discharge stroke of the actuator piston. In other embodiments, the plurality of supplementary orifices  152 B may partially overlap axially with the plurality of first orifices  152 A. The plurality of supplemental orifices  152 B dampen the driving of the actuator piston  126  due to the volume of work oil in the first actuation volume  144  that slowly vents through the plurality of supplemental orifices  152 B during driving of the actuator piston. As the actuator piston continues its stroke past the first plurality of orifices  152 A, the actuator piston  126  progressively obstructs the plurality of supplementary orifices  152 B. As the obstruction increases, the damping force increases due to work oil that remains in the first actuation volume  144  having access to less available venting area. 
     In some embodiments, the second orifices  152 B are smaller than the first orifices  152 A, which smaller diameter provides a relatively higher damping force due to less available venting area compared to an equal number of first orifices. When arranged as shown in  FIG.  11   , as the actuator piston  126  completes its stroke from right to left, the damping mechanism  105  provides an increasing damping force configuration due to several factors: progressively obstructing the first plurality of orifices  152 A, completely blocking the first plurality of orifices  152 A, the remaining second plurality of orifices  152 B being relatively smaller, progressively obstructing the second plurality of orifices  152 B, and finally completely blocking the second plurality of orifices  152 B. Accordingly, the damping mechanism  105  increases the damping force against the actuator piston  126  as it nears the end of its stroke. 
     Embodiments of the first and second orifices  152 A-B and their associated porting (including the plurality of ports  147 ) may have various shapes, sizes, and orientations. In embodiments one or both of the first and second orifices  152 A-B are circular, though other embodiments may be elongated in the direction of actuator piston driving, for example oval shaped with a long axis parallel to the direction of actuator piston  126  driving. In some embodiments, the first and second orifices  152 A-B are formed in one or more surfaces of the drive housing  114  oriented at a non-parallel angle relative to the actuator piston axis  208 . In embodiments, the orifices  152 A-B are formed in surfaces extending substantially perpendicular to the actuator piston axis  208 . In some embodiments, the first and supplemental orifices  152 A-B extend radially away from the drive cavity  116  and the actuator piston  126 . 
     In certain embodiments, 24 orifices  152  are provided in the drive cavity  116 , with 12 orifices in the first actuation volume  144  and 12 orifices in the second actuation volume  146 . Similarly, in embodiments, up to 24 first orifices  152 A and 24 second orifices  152 B are formed in the drive cavity  116  or more generally, in embodiments the number of orifices  152  may be any number from 1-48 orifices. In other embodiments, the number of orifices  152  may be greater or smaller, such as each actuation volume  144 ,  146  having up to 100 or up to 200 orifices or more. In embodiments, additional layers of orifices  152  may be included. The number of orifices may be different in different layers, for example the number of first orifices  152 A may be different than the number of second orifices  152 B. 
     Referring to  FIGS.  7 - 8   , in embodiments the drive housing  114  comprises a slight annular gap  151  between the actuator piston  126  and the drive cavity  116  and extending around an outer surface of the actuator piston (e.g., the circumferential outer surface in the illustrated embodiment). The annular gap  151  is in fluid communication with both the actuation volume  144  and the second actuation volume  146  and, in some embodiments, is configured to dampen the driving of the actuator piston  126  throughout the piston stroke in either direction by maintaining a small volume of work oil that is not in direct communication with any of the plurality of ports  147 . Accordingly, in certain embodiments, the annular gap  151  is positioned to dampen the driving of the actuator piston  126  after the orifices  152  are obstructed by the actuator piston. When the orifices  152  are fully closed leaving the plurality of ports  147  unable to vent, the relatively small annular recess  151  provides a small amount of flow area and acts like a fixed orifice during final damping. As shown in  FIG.  8   , as the actuator piston  126  strokes to the left, work oil leaves the actuation volume  144  via the orifices  152 . In certain embodiments, the circumferentially-arranged orifices  152  are fully closed off at the end of stroke; in other embodiments the orifices are fully closed off slightly before. The dampening capability of the annular gap  151  is at least in part due to compressibility of the work oil. 
     Referring to  FIGS.  9 - 11   , in embodiments, additional or final damping is provided by venting work oil through the first opening  154  and the internal porting  127 A,  127 B of the actuator piston  126 . As shown, the internal porting  127 A is in fluid communication with the plurality of ports  147 , in particular first proximal port  148 C and (indirectly) first distal port  148 A. In embodiments, the first and second proximal ports  148 C,  148 D are low-pressure ports supplying the low-pressure rail  130  and comprising a check valve preventing a vent flow out of the drive cavity. A landing orifice  107  connects the first proximal port  148 C to the first distal port  148 A, and vented work oil from the internal porting  127  flows out through the first distal port  148 A to a pressurized circuit, accumulator, or the reservoir  230 . 
     The landing orifice  107  is configured (e.g., sized) to provide desired deceleration performance of the actuator piston  126  at the end of its stroke in landing against the hard stop  106  with requisite velocity for intensifying process gas without excessive velocity that may damage components or otherwise inhibit operation. In some embodiments, the first opening  154  and the internal porting  127  vent work oil from the accumulation volume  144  throughout the stroke of the actuator piston; this venting through the first opening may occur during and/or after venting through the first and second plurality of orifices  152 A,  152 B. In certain embodiments, the internal porting  127  is configured to vent work oil after the first and second orifices  152 A-B have been completely blocked. In embodiments, the landing orifice  107  is configured to vent only after the first and second orifices  152 A-B by comprising a check valve (not shown) with a threshold pressure set at a relatively high pressure that is only achieved after the first and second orifices  152 A-B have been completely blocked. Similarly at the opposite end of the actuator piston  126 , in embodiments the internal porting  127 B is in fluid communication with the plurality of ports  147  and an additional landing orifice  107 . 
     In some embodiments, aspects of the damping mechanism  105  may be customized or tuned to increase or decrease the damping force against the actuator piston  126 . In some embodiments, the landing orifice  107  and/or one or more of the orifices  152  may comprise a removable orifice (not shown) that can be exchanged for orifices of different size or flowrate. One or more orifices  152  may comprise or a removable plug (not shown) to block one or more of the orifices by switching out. In embodiments, one or more of the annular recess  151  or the first and second plurality of orifices  152  are configured to be removable from the drive housing  114  individually or as a ring of orifices. In embodiments, this customization and tuning may optimize performance of the same compressor module  100  in different use cases (e.g., in different stack and staging arrangements or for different process gas output pressures) or in different environments (e.g., in different elevations or climates). In certain embodiments, this customization and tuning may optimize performance of the same drive housing  114  and drive cavity  116  for different pressure ratings of the compressor heads  31 ,  51 . The first and second proximal ports  148 C,  148 D may be operatively connected to the low-pressure circuit  130  or the medium-pressure circuit  132 . 
     In some embodiments, the drive housing  114  further comprises a removable sleeve insert  115  mounted internally to define the drive cavity  116  and may comprise the plurality of orifices  152 ,  152 A,  152 B and the first and second radial ports  153 ,  155  in whole or in part, along with other components of the drive housing  114 . The sleeve insert  115  is subjected to significant loads and wear forces due to the motion of the actuator piston  126  and the pressurization of the drive cavity  116 . Therefore, the sleeve insert  115  is removable for replacement after wearing down without requiring replacement of the whole drive housing  114 . In embodiments and as illustrated in  FIGS.  6 - 11   , the sleeve insert  115  comprises one or more annuli  149 A-D in fluid communication with the drive cavity  116  and a respective one or more of the plurality of ports  147 . In certain embodiments, a first and second distal annulus  149 A,  149 B are arranged respectively with the first and second distal ports  148 A,  148 B and, similarly, a first and second proximal annulus  149 C,  149 D are arranged respectively with the first and second proximal ports  148 C,  148 D. In other embodiments not illustrated, one or more of the plurality of ports  147  does not have a corresponding annulus and instead extends through the drive housing  114  and the sleeve insert  115  directly to the first or second actuation volume  144 ,  146 . 
     Each of the annuli  149 A-D extends partially or completely around the actuator piston  126  and operatively couple together the multiple discrete ports that constitute the respective port and the multiple orifices that constitute the plurality of orifices  152 ,  152 A,  152 B. For example in  FIGS.  10 - 11   , the first distal annulus  149 A connects the multiple first distal ports  148 A with both of the first plurality of orifices  152 A (corresponding radial ports not shown) and the second plurality of orifices  152 B (through first radial ports  153 ) and the first proximal annulus  149 C connects the multiple first proximal ports  148 C with the first internal porting  127 A. In the example of  FIG.  7   , the first distal annulus  149 A connects each of the multiple first distal ports  148 A arranged around the actuator piston  126  with the manifold port  117 . 
     In embodiments with first and second compressor heads  31 ,  51 , during the discharge cycle of the second compressor head  51 , the drive cavity  116  is configured to similarly dampen the driving of the actuator piston  126 . A volume of work oil in the second actuation volume  146  provides damping force and vents through the plurality of orifices  152 ,  152 A during driving of the actuator piston  126 . The plurality of orifices  152 ,  152 A are open to the second actuation volume  146  when the driving of the actuator piston  126  begins, and the plurality of orifices are progressively covered by the actuator piston during the driving. Covering the plurality of orifices increases the damping force of work oil remaining in the second actuation volume  146  acting against the second side  145  of the actuator piston  126 . In embodiments, the internal porting  127 B and corresponding landing orifice  107  provide damping and control the final velocity of the actuator piston  126  impacting the respective hard stop  106 . 
     It will be appreciated that in embodiments, any of the ports and orifices that provide damping can be reversible and configured to provide an actuation supply of pressurized work oil for the actuator piston  126 , for example in a discharge cycle of the second compressor head  51 . In embodiments, the actuation supply is provided sequentially in a reverse order from the damping sequence above. Accordingly, for the discharge cycle of the second compressor head  51  in  FIGS.  9 - 11   , the actuation supply of work oil to the first actuation volume  144 , work oil (e.g., low-pressure work oil from the low-pressure circuit  130 ) begins from the first proximal port  148 C and then through the first internal porting  127 A of the actuator piston  126  feeding to the first opening  154 . Subsequently, the second plurality of orifices  152 B is configured to provide additional actuation supply of pressurized work oil via the first distal port  148 A. Finally, the first plurality of orifices  152 A is configured to provide additional actuation supply of pressurized work oil via the first distal port  148 A. In embodiments, this sequential actuation supply provides one or more of the plurality of pressure circuits  120 . In some embodiments, one or more ports of the plurality of ports  147  is configured to further supplement the actuation supply of work oil provided by another port. In certain embodiments the first proximal port  148 C comprises a bypass check valve (not shown) to provide supplemental flow in addition to the first distal port  148 A, which may be configured to avoid cavitation of work oil in the first actuation volume  144  as the actuator piston  126  moves quickly. As discussed above, the initial movement of the actuator piston  126  may be aided by additional forces such as the return of the diaphragm  5  of the first compressor head  31 . 
     The damping mechanism  105  may be advantageous for embodiments of the compressor module  100  with a short stroke of the actuator piston  126 , for example about 1″ or 2.5″, or more generally a stroke below about 7.5″. With shorter travel distance, the peak speed of the actuator piston  126  may be lower than with a relatively longer stroke, and deceleration at the end of the stroke may achieve low impact velocities. In certain embodiments, compressor heads  31 ,  51  that are configured for higher pressures comprise a stroke of about 1″ (e.g., about 7,500 psi and above, including embodiments comprising about 7,500 psi and about 15,000 psi and ranges therebetween), whereas compressor heads configured for relatively lower pressure comprise a stroke of about 2.5″ (e.g., about 5,000 psi and below, specific embodiments comprising 1,000 psi; 2,000 psi; 5,000 psi; and ranges therebetween). 
     As discussed above, in certain embodiments of the damping mechanism  105 , the actuator piston  126  closes off circumferential primary supply ports (such as the first and second distal ports  148 A,  148 B) as it reaches the hard stop  106 . Dynamic computer simulations of such embodiments of the damping mechanism  105  show that, for embodiments of the actuator pistons  126  with 1″ stroke, the impact velocities are less than about 0.2 m/s for nominal conditions and less than about 0.7 m/s for avoiding a failure condition. The simulations assumed nominal radial clearances in a range between about 0.0025-0.010 inches between the actuator piston  126  and inner walls of the drive cavity  116 , although smaller clearances are contemplated for other embodiments. For the actuator pistons  126  with 2.5″ stroke and higher pressure supply, the impact velocities were less than about 0.3 m/s for nominal conditions and less than about 1.5 m/s for avoiding a failure condition. It will be appreciated that other embodiments may comprise a broader range of values for parameters such as the stroke distance of the actuator piston  126 , landing velocity, and output pressure of the compressor heads  31 ,  51 . 
     Main Stage Valve 
     In certain embodiments, the high-throughput compressor system  200  comprises one or more main stage valves  250  (“MSV  250 ”) to control the hydraulic drive  110 , in particular the flow of work oil to and from the drive cavity  116 . 
     The work oil drives and damps the actuator piston  126 , therefore the MSV(s)  250  control the timing (cycle time, travel time) of the actuator piston  126 . As detailed above, the actuator piston  126  discharge stroke travel distance may be in a range of about 0.5-3 inches with a travel time of less than 100 milliseconds, other embodiments may range from 0.5-7 inches or more. Accordingly, the timing of the MSV(s)  250  in supplying and venting work oil from one or more of the plurality of pressure circuits  120  must correspond to these parameters. 
     In some embodiments, the MSV(s)  250  control the interface of the HPU  118  and of the one or more pressure circuits  120  with the hydraulic actuator  112 , such interface including both the supply and vent of work oil for the drive cavity  116 . In other words, the MSV(s) control a pressurized hydraulic supply of work oil for operating the hydraulic actuator  112  and the MSV(s) may control at least some venting of work oil from the drive cavity  116 . In embodiments, the MSV  250  is an actively-controlled valve. In the illustrated embodiment, the MSV  250  is a three-way valve as shown in  FIGS.  18 A  (vent stage) and  18 B (supply stage). 
     Referring to  FIGS.  18 A-B , in embodiments, The MSV  250  for controlling a diaphragm compressor system  100  comprises a valve body  260  comprising a first end  262  and a second end  264 , a pilot port  266  proximate the first end, a supply port  268 , a first vent port  270 , a second vent port  271 , and a cylinder port  272 . The MSV  250  comprises a pin subassembly  274  for reciprocating in the valve body  260 , the pin comprising a spool  276 , a pilot pin  278  proximate the second end  264 , and a return pin  280  proximate the first end  262 . In certain embodiments, the MSV  250  is mounted to the valve manifold  244  or the drive housing  114 . In some embodiments, each of the vent port  270  and the cylinder port  272  are in fluid communication with the drive cavity  116  and the supply port  268  operatively coupled to the hydraulic power unit  118 . In embodiments, the position of the pin subassembly  274  selectively blocks one or both of the vent port  270  and the cylinder port  272  to control the flow of work oil between the drive cavity  116  and the MSV  250  along with controlling the flow of work oil between the MSV  250  and any other component(s) attached to the MSV. It will be appreciated that any of the ports may be one or more ports in some embodiments, and in the illustrated embodiment each of the pilot port  266 , supply port  268 , the first vent port  270 , and the cylinder port  272  is a plurality of ports arranged annularly about the pin subassembly  274 . 
       FIG.  18 A  shows a vent position of the MSV  250  configured for allowing an outflow of work oil from the drive cavity  116 , e.g. from work oil in the first or second actuation volume  144 ,  146  when being compressed by the actuator piston  126  and vented through the plurality of orifices  152 . The pin subassembly  274  is configured to move axially to the vent position with the cylinder port  272  in fluid communication with the first vent port  270 , such that work oil from the drive cavity  116  flows to the cylinder port  272 , through the valve body  260 , and out through the first vent port  270 . In embodiments, the hydraulic drive  110  further comprises a recovered oil accumulator  136 D operatively coupled to the first vent port  270  of the MSV  250  for storing and recycling this work oil in subsequent cycles. In other embodiments, the first vent port  270  is operatively connected to a reservoir  230  of the hydraulic drive  110 . 
     In embodiments, the MSV  250  comprises one or more vent orifices  282 A,  282 B configured to vent work oil out of the MSV and dampen motion of the pin subassembly  274  when moving into the vent position. The vent orifices  282 A,  282 B are arranged annularly around the pin subassembly  274 . Similar to the damping mechanism  105  of the actuator piston  126 , the vent orifices  282 A in the MSV  250  are configured to be progressively obstructed as the pin subassembly  274  moves axially, increasing the damping force as the pin reaches the end of its stroke. Accordingly, the pin subassembly  274  moves quickly to the vent position without any hard impact or bounce. In the illustrated embodiment, vent orifices  282 B are not configured to be obstructed, but in other embodiments these or additional layers of orifices (not shown) may be configured to be obstructed in addition to the vent orifices  282 A. 
       FIG.  18 B  shows a supply position of the MSV  250  configured to supply work oil to the hydraulic actuator  112 . The MSV  250  is configured to selectively move to the supply position to operatively connect the HPU  118  to the drive cavity  116  during the discharge cycle of the first or second compressor head  31 ,  51 . In embodiments, the MSV  250  supply position is configured to control a discharge stroke of the compressor head  31 , wherein the MSV  250  connects one of the pressure circuits  120  to the second actuation volume  146  of the drive cavity  116  to supply pressurized work oil to the second side  145  of the actuator piston  126  to drive the actuator piston and consequently drive the diaphragm piston  3  toward the diaphragm  5  of the first compressor head  31 . This connection in embodiments is from the high-pressure circuit  134 , medium-pressure circuit  132 , or low-pressure circuit  130 . The pin subassembly  274  is configured to move axially to the supply position with the supply port  268  in fluid communication with the cylinder port  272 , such that work oil flows from the HPU  118  or a pressure circuit enters through the supply port  268 , passes through the valve body  260 , and exits through the cylinder port  272 . An end orifice  284  is located proximate the first end  262  of the MSV  250 . The pilot port  266  and the end orifice  284  configured to vent work oil out of the MSV  250  and dampen motion of the pin subassembly  274  when moving into the supply position in a similar manner to the vent orifices  282 A,  282 B for the vent position. 
     In embodiments, during a suction cycle of the first compressor head  31 , the MSV  250  is configured to move to the vent position ( FIG.  18 A ) to connect the drive cavity  116  of the drive housing  114  to the first vent port  270  of the main stage valve  250 , and the hydraulic drive  110  vents work oil from the second actuation volume  146  to the main stage valve  250  through the vent port  270 . 
     In embodiments, one or more of the pilot port  266  and the vent orifice  282 A-B comprises a plurality of rows of orifices that are axially spaced. In certain embodiments, this plurality of rows comprises a row of relatively larger orifices proximate the spool  276  and a row of smaller orifices proximate the respective first or second end  262 ,  264 . In some embodiments, the pilot port  266  comprises a ring or layer of removable orifices, or plugs for certain orifices, for fine tuning of the damping performance. Such tuning may be necessary for implementing the MSV  250  with different pressure ratings of compressor heads  31 ,  51 , for utilizing different types of work oil, or for operating at different temperature ranges. 
     In embodiments, the low-pressure circuit  130  of the hydraulic drive  110  further comprises a recovered oil accumulator  136 D operatively coupled to the first vent port  270  of the main stage valve  250 . In such embodiments, in the vent position, the main stage valve  250  is configured to supply oil from the drive cavity  116  to the cylinder port  272 , through the valve body  260 , and exiting the first vent port  270  to the recovered oil accumulator. In some embodiments, the low-pressure circuit  130  comprises the recovered oil accumulator  136 D. In certain embodiments, a passive valve  131  ( FIG.  17   ) is operatively connected to the recovered oil accumulator  136 D and the drive cavity  116  downstream of the main stage valve  250 . During the suction cycle of the first compressor head, the passive valve  131  is configured to supply oil from the recovered oil accumulator  136 D to the drive cavity  116 . In embodiments, this supply from the recovered oil accumulator  136 D may occur at one or more times, for example during a low-pressure stage of tank fill or during a beginning portion of each actuator  126  stroke. 
     In some embodiments, the MSV  250  further comprises a pilot valve  290  configured to selectively actuate the pin subassembly  274  of the MSV  250 . The pilot valve  290  controls a supply of pilot fluid (e.g., work oil at a pilot pressure) to the MSV  250  to move the pin subassembly  274 . The pilot valve  290  in the illustrated embodiment is mounted in the second end  264  of the valve body  260  and is a multi-stage valve comprising a spool and two coils. In some embodiments, the hydraulic drive  110  further comprises a pilot pressure circuit  138  and a pilot pressure accumulator  136 E operatively coupled to the pilot valve. The pilot pressure accumulator  136 E in embodiments is charged in various ways such as a separate hydraulic unit, the HPU  118 , or recovered intensified work oil vented from one or more MSVs  250 . In some embodiments, the pilot valve  290  is a three-way valve or two two-way valves. In other embodiments, a different actuator or valve (not shown, for example a spool valve with one coil and one spring return, a piezo actuator, or servomotor) is implemented with the MSV  250  to selectively actuate the pin subassembly  274  of the MSV  250  to the supply position 
     In embodiments, the pilot pressure circuit  138  is also operatively coupled to the pilot port  266  at the first end  262  of the valve body  260 . In some embodiments, the pin subassembly  274  of the MSV  250  has a larger area proximate the pilot valve  290  than proximate the pilot port  266 . Therefore, when pilot pressure is supplied to the pilot pin  278  through the pilot valve  290  and the pilot port  266 , the pin subassembly  274  is configured to move to the supply position. In embodiments, the MSV  250  comprises a return spring  286  configured to bias the pin subassembly  274  toward the vent position when pressure is not supplied to the pilot valve  290 . As shown in  FIG.  15   , each of multiple MSVs  250 A-D may comprise a respective pilot valve  290 . 
     As noted above, some embodiments of the high-throughput compressor system  200  and/or the individual compressor modules  100  comprise multiple MSVs  250 . Referring also to  FIGS.  15 - 17   , in embodiments a first MSV  250 A is operatively coupled to the first actuation volume  144  of the drive cavity  116 . A second MSV  250 B is substantially similar to the first MSV  250 A, and the second MSV  250 B is mounted to the drive housing  114  with each of the vent port  270  and the cylinder port  272  in fluid communication with the second actuation volume  146  of the drive cavity  116  and the supply port  268  operatively coupled to the hydraulic power unit  118 . The first MSV  250 A is configured to selectively move to the supply position to connect the high-pressure circuit  134  to the second actuation volume  146 , while the second MSV  250 B is configured to selectively move to the respective supply position to connect the medium-pressure circuit  134  to the second actuation volume  146 . 
     In other embodiments, four MSVs  250 A-D are provided with a first MSV  250 A connecting the high-pressure circuit  134  to the first actuation volume  144 , a second MSV  250 B connecting the high-pressure circuit  134  to the second actuation volume  146 , a third MSV  250 C connecting the medium-pressure circuit  132  to the first actuation volume  144 , and a fourth MSV  250 D connecting the medium-pressure circuit  132  to the second actuation volume  146 . 
     In embodiments one or more of the MSVs  250 A-D vents to the recovered oil accumulator  136 D. In the illustrated embodiment, the third and fourth MSVs  250 C,  250 D are each configured to vent from the respective first or second actuation volume  144 ,  146  through the respective first vent port  170  to the accumulator  136 D. 
     Referring to  FIG.  20   , the valve manifold  244  is illustrated with some of the corresponding hydraulic components and internal plumbing. The MSV  250  is mounted in a valve mount  292  that is plumbed to the operative ports of the MSV  250 . The low-pressure circuit  130  is collected from several sources including each MSV  250  along with vented return from the hydraulic drive  110 , these sources lead to the recovered oil accumulator  136 D. The high-pressure circuit  134  is ported externally from the HPU  118  (not shown, see also  FIG.  29   ). Embodiments comprising the medium-pressure circuit  132  are similarly supplied by another HPU  118 . It will be appreciated that the valve manifold  244  in embodiments may be operatively connected to multiple compressor modules  100 , reducing the overall footprint of the diaphragm compressor system  200 . In certain such embodiments, each MSV  250  is configured and operatively connected to the multiple compressor modules  100 . In other such embodiments, the valve manifold  244  comprises one or more additional MSVs for the additional compressor module(s). 
     In some embodiments, the location and orientation of ports in the MSV  250  are selected to fit in the valve housing  244  ( FIG.  20   ) while accommodating nearby components. In certain embodiments, pilot port  266 , the supply port  268 , and/or the first and second vent ports  270 ,  271  are arranged to allow work oil to enter and exit on a same side of the valve mount  292  ( FIG.  20   ) that is opposite from the pilot pressure circuit  138  and other control components. Embodiments of the present disclosure provide sufficiently large flow areas through the ports to result in minimal or substantially no pressure drop of pressurized fluid passing through the MSV  250 . In some embodiments, the MSV  250  is configured to accommodate pressures up to 5,000 psi and provides an effective flow area (CdA, i.e., discharge coefficient×area) of about 300 mm 2 . In embodiments, the MSV  250  provides a CdA of about 275-325 mm 2 , about 250-350 mm 2 , about 200-400 mm 2 , about 100-500 mm 2 , at least 200 mm 2 , at least 250 mm 2 , or at least 300 mm 2 . In embodiments, the MSV  250  is configured to accommodate pressures up to 15,000 psi. 
     In other embodiments, other valve types are employed in addition to or in lieu of the MSV  250 , including poppet, spool, directional, proportional and servo valves, among others. Different types of valves could be used as MSV  250  to operate the system differently. In some embodiments, proportional valves control the flow into the system with a fixed supply pressure. In this way the valve could be used to speed up or slow down the travel of the hydraulic drive actuator to fit a desired profile or to reduce the velocity of the actuator  112  as it nears top dead center or bottom dead center. 
     In other embodiments, digital or on/off valves allow full flow to be supplied to (or vented from) the MSV  250  with a fixed flow area. As these valves open to the pressurized supply of work oil, the maximum flow area is exposed and allows full flow into the MSV  250  as dictated by the differential pressure across the valve. These valves are closed to shut off flow to the hydraulic actuator  112  for embodiments as a two-way valve. These valves can also vent the hydraulic actuator  112  for embodiments as a three-way valve. In still other embodiments, a variation of the digital on/off valve has multiple outlet ports that could be opened in series to allow flow to variable areas within the hydraulic drive. In this valve, the internal spool moves only a portion of its travel distance to open up flow to a single outlet port, then as the spool continues its travel additional outlet ports are opened. Operation of the digital valves can be achieved in several ways. In embodiments, the digital MSVs  250  are operated with a solenoid to drive the valve. In other embodiments, the digital MSVs  250  are operated with a set of two-way pilot valves to control the supply of pilot fluid to drive the valve spool. In other embodiments, the digital MSVs  250  are operated with a single three-way pilot valve to control the supply of pilot fluid to drive the valve spool. It will be appreciated that in embodiments, the MSVs  250  can be combinations of one or more of the above valve types. 
     Active Oil Injection System 
     In some embodiments, the diaphragm compressor  1  employs a hydraulic injection pump system  10 . The hydraulic injection pump system  10  comprises a pump  12 , at least one oil check valve  45  and a fixed setting oil relief valve  14  as illustrated in  FIG.  33   . Other embodiments discussed below replace the fixed setting oil relief valve with a variable pressure relief valve  52  (“VPRV  52 ”). The injection pump system  10  primary function is to maintain the required oil volume between the high-pressure oil piston  3  and diaphragm set  5 . During the compressor  1  (e.g., compressor head  31  or  51 ) suction stroke, a fixed volume of work oil is injected into the work oil region  35  of the compressor  1 . This ensures a sufficient volume of oil is injected during each suction stroke to ensure the oil volume is maintained for proper compressor  1  performance. 
     In certain embodiments the oil volume between the diaphragm piston  3  and diaphragm  5  is impacted by two modes of oil loss. The first mode of oil loss is annular leakage past the diaphragm piston  3  (also referred to as a high-pressure oil piston) back to the drive housing  114  or an oil reservoir. This annular leakage may be most significant on high pressure compressors  1  operating above 5,000 psi. In some embodiments, the annular leakage varies during operation of the diaphragm compressor  1 . 
     The second mode of oil loss is defined as “overpump” which is hydraulic flow over the oil relief valve  14  that occurs every cycle during normal compressor  1  operation. The injector pump system  10  is designed and operated to maintain an “overpump” condition through the relief valve  14  ensuring the diaphragms  5  are sweeping the entire compressor cavity  15  (i.e., completely or substantially discharging process gas from the process gas region  36 ) thereby maximizing volumetric efficiency of the compressor  1 . Embodiments of the present disclosure comprise an injection pump system  10  that is actively controlled, referred to as an active oil injection system (“AOIS”)  30  as further discussed below. 
     Some embodiments of the injection systems  10  are mechanically adjustable by a user to vary the injector pump&#39;s  12  volumetric flow rate into the compressor  1 . However, this requires manual observations and adjustment. An incorrect volumetric displacement from the injection pump system  10  that does not sufficiently account for oil losses can lead to various machine failures. 
     In certain embodiments, the hydraulic relief valve  14  has a manually adjustable relief setting. These oil relief valves are set to a fixed oil relief pressure setting that is higher than the maximum process gas pressure. The maximum process gas pressure is the maximum expected pressure of the process gas for any particular use case. This elevated relief setting allows the diaphragm  5  to contact the process gas head support plate  6  firmly before any work oil flows over the relief valve  14 , thus, assuring a complete sweep of the entire volume of the head cavity  15  at the highest expected pressure of the process gas. When the diaphragm reaches the top of the head cavity  15 , the diaphragm piston  3  still has a pressure below the setting of the relief valve  14 . During this period, the work oil in the work oil region  35  compresses further and the hydraulic pressure rises above the compressor gas discharge pressure until it reaches the setting of the oil relief valve  14 . At this point, the relief valve  14  opens and oil, in the amount of the injection pump displacement (i.e., injection volume) less the annular leakage in the system, is displaced over the oil relief valve  14 . This oil flow out of the relief valve  14  is defined as overpump. Because the annular leakage may vary during operation, in some embodiments the injection volume does not correlate or loosely correlates to the volume of overpump flow through the relief valve  14 . In other embodiments, the injection volume corresponds or correlates to the volume overpump flow (for example, when the annular leakage has only minor variation, the annular leakage is variably estimated for different operating conditions, or the annular leakage is measured or otherwise detected). 
     Certain embodiments of the present invention include an active oil injection system  30  (“AOIS”) in a diaphragm compressor  1 . The feedback and control of the AOIS  30  allow the compressor system  100  to minimize any excess energy used while ensuring the complete sweep of the diaphragm  5  discussed above. 
     In certain embodiments, the compressor  1  forms a hydraulic circuit  50  connecting the outlet  34  of the work oil head support plate  8  to the inlet  33  of the work oil head support plate  8 . In those embodiments, the hydraulic circuit may also include an oil reservoir  38  configured to collect overpumped work oil from the work oil region  35  via the outlet  34  of the work oil head support plate  8 . By forming a hydraulic circuit, oil is circulated from the oil reservoir  38 , through the inlet  33  and into the work oil region, and then out the outlet  34  and back into the oil reservoir  38 . In another sense, work oil that exits the outlet  34  and passes through the oil relief valve  14  constitutes the overpumped work oil from the compressor  1 . 
     In other embodiments, the hydraulic circuit also includes an AOIS  30  including a hydraulic accumulator  39  configured to provide a supply of supplemental work oil to the inlet  33  of the work oil head support plate  8 . In certain embodiments, the hydraulic accumulator  39  may be a hydraulic volume or any style of hydraulic accumulator  39  such as a bladder, piston, or diaphragm gas over fluid style hydraulic accumulator  39 . In still further embodiments, the AOIS includes an AOIS pump  40  in communication with the hydraulic accumulator  39 , the AOIS pump  40  configured to produce a variable volumetric displacement of the supplemental work oil from the oil reservoir  38  to the hydraulic accumulator  39  or directly to the inlet  33 . As used herein, variable volumetric displacement means that the AOIS  30  can provide a variable volumetric flow (i.e. injection quantities of supplemental work oil via the pump  40 ) and/or an independently variable speed (i.e., flow rate via the motor  41 ), to the work oil region  35  depending on the particular process conditions of the compressor  1  (e.g., compressor head  31 ). This allows for variable injection quantities during the compressor&#39;s  1  operation to maintain the compressor&#39;s  1  oil volume most efficiently within the compressor  1 , and particularly the work oil region  35 . In certain embodiments, the AOIS  30  includes the AOIS pump  40  operatively coupled to the hydraulic accumulator  39 , and a motor  41  configured to power the AOIS pump  40  independently from the hydraulic drive  110 . In other words, the speed and control of the motor  41  is completely independent from, and not mechanically linked to, the hydraulic drive  110  that powers the diaphragm piston  3 . 
     In certain embodiments, the AOIS  30  utilizes the existing pressure dynamics within the compressor  1  to satisfy the hydraulic flow requirements into the compressor  1 , and particularly into the work oil region  35 . As the compressor  1  transitions through its suction and discharge cycles, the AOIS pump  40  charges and discharges the hydraulic accumulator  39 . During the compressor&#39;s  1  suction stroke, this lower pressure condition within the compressor  1 , including the work oil region  35 , creates a positive pressure differential between the hydraulic accumulator  39  and the oil within the compressor head  31 , and particularly in the work oil region  35 . During this suction condition, hydraulic flow goes through the oil inlet check valves  45  and through inlet  33  into the work oil region  35  satisfying the injection event. During this time, the pump  40  may be continuously pumping into the hydraulic accumulator. During this discharge stroke, the hydraulic pressure within work oil region  35  is greater than the pressure in the hydraulic accumulator  39  therefore there is no flow from the hydraulic accumulator  39  into the compressor. At least one check valve  45 , and in some embodiments at least two check valves  45 , prevent backflow from the work oil region  35  into the hydraulic accumulator  39  and beyond. During this this condition, the hydraulic flow from the AOIS pump  40  pressurizes the hydraulic accumulator  39  in preparation for the next injection event. 
     Further embodiments include a variable pressure relieve valve (VPRV)  52 , which includes a pressure relief mechanism  42  operatively coupled to the work oil region  35  of the diaphragm cavity  15 , the pressure relief mechanism  42  including a pressure relief valve  43  in communication with the outlet  34  of the work oil head support plate  8  and configured to relieve an outlet volume of the pressurized work oil from the work oil region  35 . In these embodiments, the pressure relief valve  43  includes a hydraulic relief setting corresponding to an overpump target condition of the pressurized work oil relative to the process gas discharge pressure. In some embodiments, the overpump target condition corresponds to a maximum process gas discharge pressure. In other words, the overpump target condition corresponds to a maximum process gas discharge pressure that the compressor head  31  is configured to operate at, so that the process gas region  36  is configured to be completely evacuated by the diaphragm  5  at maximum gas discharge pressure. 
     In certain embodiments, during an oil relief event during the discharge cycle, the relief valve  43  opens and oil, in the amount of the injection volume per revolution less the annular leakage in the system, is displaced over the oil relief valve  14 , defined as overpump. During this time, the hydraulic flow from the AOIS pump  40  pressurizes the hydraulic accumulator  39  in preparation for the next injection event during the next suction cycle. 
     However, in certain embodiments, the pressure relief valve  43  is configured to actively adjust the hydraulic relief setting of the pressure relief valve to correspond to an overpump current condition. In other words, the pressure relief valve  43  is configured to adjust the hydraulic relief setting up or down corresponding to a relative increase or decrease in gas discharge pressure. This prevents the compressor head  31  from experiencing more overpump than necessary to completely evacuate the process gas region  36  by the diaphragm  5  under conditions with a gas discharge pressure less than the maximum gas discharge pressure. Adjustability of the hydraulic relief setting may enable longer machine life expectancy and better system efficiency due to lower cyclic stresses and lower alternating loads during the compressor&#39;s  1  discharge and suction cycles. 
     Certain embodiments of the AOIS  30  include an injector pump  40  and hydraulic accumulator  39  without a VPRV  52 , while other embodiments include both systems. 
     In certain embodiments, the AOIS  30  includes a feedback mechanism configured to control the AOIS pump  40  to maintain the overpump target condition of the work oil region  35 . The feedback mechanism includes a measurement device  44  that provides feedback to verify the over pump condition is being met to control the injector pump system  30 . In certain embodiments, the feedback mechanism includes a first measurement device  44  operatively coupled to the diaphragm compressor  1 , the measurement device configured to detect and/or measure the overpump current condition of the intensified work oil flowing out of the outlet  34  from the work oil region  35 . In certain embodiments, the feedback mechanism is configured to adjust the volumetric displacement of the injector pump  40  to the hydraulic accumulator  39  in response to the overpump current condition. 
     Turndown ratio refers to the operational range of a device, and is defined as the ratio of the maximum capacity to minimum capacity. In certain embodiments of the AOIS  30 , the AOIS is configured to provide a large turndown ratio of supplemental work oil relative to the work oil  4  in the work oil region  35  of the compressor  31 . By separating the functions of the hydraulic drive  31  and the AOIS pump  40 , a large turndown ratio can be achieved allowing for significant adjustability of injection quantity to tightly control the amount of overpump through the relief valve  43  over a wide range of operating conditions. 
     In embodiments, the overpump target condition ranges from 0.1%-500% above a measured process gas discharge pressure. In various embodiments, the overpump target condition ranges from about 0.1%-100% above, 0.1%-50% above, 0.1%-40% above, 0.1%-30% above, 0.1%-20% above, 1%-20% above, or 1%-50% above the measured process gas discharge pressure. 
     Alternative Embodiments 
     Similar stacked compressors and related systems are also disclosed in U.S. patent application Ser. Nos. 17/840,919; 17/840,937; and 17/840,948 each filed Jun. 15, 2022, the entire contents of which are incorporated herein by reference and for all purposes. 
     Some embodiments incorporate commonality of parts and assemblies between stages  202  even though the later stages may have larger compressor heads, higher pressure rails/circuits, and the like. Specifically, such items as valve manifolds  244 , MSVs  250 , and other hydraulic components can be common for cost and simplicity purposes. Additionally, the clamping mechanism  204  may have duplicate components or similar primary components with minor deviations to accommodate adapting and mating to specific stages. 
     In other embodiments, the first and second compressor heads  31 ,  51  are driven by two separate hydraulic actuators  112  instead of a single hydraulic actuator, and the two hydraulic actuators may be configured to act in parallel or phase with each other such that the discharge and suction cycles of the first and second compressor heads  31 ,  51  occur substantially simultaneously. Although certain embodiments of the disclosed compressor modules  100  are hydraulically driven, in other embodiments, other modes of actuating the diaphragms  5  may be implemented. In embodiments, one or more compressor modules  100  may be driven by a crank-slider mechanism (not shown) or other mechanism. 
     Applicable to any embodiments disclosed herein, the terms “upward” and “downward” are used for convenience in reference to the figures for explaining examples of motion, but are not meant to be limiting. In embodiments, the diaphragm piston  3 , diaphragm  5 , and other components may move in any direction relative to each other, for example left and right, inward and outward, and the like. In embodiments, the diaphragm piston  3  may move perpendicularly or otherwise angled relative to the diaphragm  5  or relative to the actuator piston  126  or other components of the hydraulic drive  110 , so long as actuation movement of the diaphragm piston  3  pressurizes work oil against the diaphragm. In embodiments, the diaphragm piston  3  or intermediate pistons  183  may move in a direction away from or offset from the diaphragm  5 . In other words, by referring to the movement of the piston as the terms “upward” and “downward” with respect to the diaphragm  5  or the compressor head, those terms may be understood as “toward” and “away from,” respectively, or may understood as “pressurizing the work oil” and “depressurizing the work oil,” respectively, or “discharge cycle” and “suction cycle,” respectively. 
     All of the features disclosed, claimed, and incorporated by reference herein, and all of the steps of any method or process so disclosed, may be combined in any combination, except combinations where at least some of such features and/or steps are mutually exclusive. Each feature disclosed in this specification may be replaced by alternative features serving the same, equivalent or similar purpose, unless expressly stated otherwise. Thus, unless expressly stated otherwise, each feature disclosed is an example only of a generic series of equivalent or similar features. Inventive aspects of this disclosure are not restricted to the details of the foregoing embodiments, but rather extend to any novel embodiment, or any novel combination of embodiments, of the features presented in this disclosure, and to any novel embodiment, or any novel combination of embodiments, of the steps of any method or process so disclosed. 
     Although specific examples have been illustrated and described herein, it will be appreciated by those of ordinary skill in the art that any arrangement calculated to achieve the same purpose could be substituted for the specific examples disclosed. This application is intended to cover adaptations or variations of the present subject matter. Therefore, it is intended that the invention be defined by the attached claims and their legal equivalents, as well as the illustrative aspects. The above described embodiments are merely descriptive of its principles and are not to be considered limiting. Further modifications of the invention herein disclosed will occur to those skilled in the respective arts and all such modifications are deemed to be within the scope of the inventive aspects.