阅读说明：本技术 带有偏移滑块曲柄的电动隔膜泵 (Electric diaphragm pump with offset slider crank ) 是由 D.布洛克 Z.阿斯特 J.西蒙斯 S.J.L.兹迪克 A.G.布拉格斯 W.A.赛思 于 2020-03-10 设计创作，主要内容包括：一种隔膜泵,其具有能够围绕旋转轴线旋转并联接于活塞的曲轴。活塞能够在吸入冲程和排出冲程之间在活塞缸内沿着运动轴线往复地移位。联接于活塞缸的隔膜壳体至少部分地限定泵送腔室,在活塞往复运动时,流体泵送通过该泵送腔室。与活塞和连杆之间的连接相交的运动轴线可不与曲轴的旋转轴线相交,使得相对于其中运动轴线与旋转轴线相交的布置,在排出冲程期间活塞侧负载力的峰值减小,并且在吸入冲程期间活塞侧负载力的峰值增大,以便获得排出冲程和吸入冲程的活塞侧负载力的峰值之间的改善的平衡。(A diaphragm pump has a crankshaft rotatable about an axis of rotation and coupled to a piston. The piston is reciprocally displaceable within the piston cylinder along an axis of motion between a suction stroke and a discharge stroke. A diaphragm housing coupled to the piston cylinder at least partially defines a pumping chamber through which fluid is pumped as the piston reciprocates. The movement axis intersecting the connection between the piston and the connecting rod may not intersect the rotational axis of the crankshaft, so that the peak value of the piston-side load force during the discharge stroke is reduced and the peak value of the piston-side load force during the suction stroke is increased, relative to an arrangement in which the movement axis intersects the rotational axis, in order to obtain an improved balance between the peak values of the piston-side load force of the discharge stroke and the suction stroke.)

1. An apparatus, comprising:

a diaphragm pump having a plurality of diaphragms at least partially defining a plurality of pump chambers, each diaphragm defined to be structured to reciprocate between a first position and a second position to pump a fluid, the diaphragm pump comprising:

a crankcase;

a bracket configured to support the crankcase, the bracket configured to contact a horizontal support surface;

a crankshaft disposed at least partially within the crankcase, the crankshaft having a crankshaft axis oriented in a vertical direction by the bracket, the crankshaft being operatively connected to a motor such that the motor provides power to rotate the crankshaft during operation;

at least one cam configured to rotate with the crankshaft to cause the reciprocating motion of the plurality of diaphragms, wherein each of the plurality of diaphragms is oriented to reciprocate along a cylinder axis orthogonal to the crankshaft axis.

2. The apparatus of claim 1, wherein the cylinder axes of the plurality of diaphragms are vertically displaced relative to each other.

3. The apparatus of claim 1, wherein the diaphragm pump comprises at least three diaphragms and at least three pumping chambers, and wherein the at least three pumping chambers are evenly distributed about the crankshaft axis such that angles formed between adjacent crankshaft axes are substantially the same.

4. The apparatus of claim 3, wherein each of the at least three diaphragms is reciprocated by a single cam.

5. The apparatus of claim 1, wherein the motor is an electric motor having a rotatable rotor in power communication with the crankshaft and rotatable about a rotor axis parallel to the crankshaft axis.

6. The apparatus of claim 5, wherein the motor is configured to be driven to reduce a speed of the crankshaft and maintain torque upon an occurrence of a flow interruption event.

7. The apparatus of claim 1, wherein the bracket comprises a plurality of legs extending from the crankcase.

8. The apparatus of claim 1, wherein each of the cylinder axes is offset from the crankshaft axis, and wherein the diaphragm pump further comprises a cylinder for each of the plurality of diaphragms, each cylinder being arranged along the cylinder axis, and a piston disposed within the cylinder, the piston being connected to the crankshaft by a connecting rod.

9. The apparatus of claim 8, wherein each of the pistons is structured to move the respective diaphragm directly between the first position and the second position.

10. The apparatus of claim 9, wherein the reciprocal displacement of each of the pistons along the cylinder axis is guided by rolling element bearings between the piston and the cylinder.

11. The apparatus of claim 9, wherein a seal is disposed between the piston and the cylinder to prevent lubricant from reaching a receiving cavity opposite the pumping chamber relative to the diaphragm.

12. The apparatus of claim 11, wherein the seal comprises a lubricant facing seal and a containment cavity facing seal, wherein the containment cavity facing seal is a bellows seal.

13. The apparatus of claim 8, wherein the diaphragms are oriented such that the arcuate shape of the annular flexible portion of each diaphragm is disposed in a direction generally away from the pumping chamber.

14. The apparatus of claim 8, wherein each of the pistons comprises a support band positioned circumferentially around at least a portion of the respective piston.

15. A method, comprising:

inserting a crankshaft at least partially into a crankcase of a diaphragm pump, the crankshaft having at least one cam and a crankshaft axis;

coupling an electric motor to the crankcase, the electric motor in power communication with the crankshaft;

coupling a bracket to the crankcase such that the crankshaft axis is vertically oriented when the crankcase is supported by the bracket;

orienting a cylinder axis of each of a plurality of diaphragm pump cylinders with respect to the crankshaft, each of the respective cylinder axes forming a right angle with the crankshaft axis when viewed from a side transverse to the respective cylinder axis and the crankshaft axis; and

attaching the plurality of diaphragm pump cylinders to the crankcase, the diaphragm pump cylinders being equally angularly spaced about the crankshaft axis.

16. The method of claim 15, further comprising orienting the bracket such that respective cylinder axes of the plurality of diaphragm pump cylinders are horizontal.

17. The method of claim 15, further comprising positioning a piston within each of the respective plurality of diaphragm pump cylinders and connecting the diaphragm in each respective cylinder to the respective piston.

18. The method of claim 17, further comprising attaching a connecting rod between each piston and the camshaft.

19. The method of claim 15, wherein orienting the cylinder axis comprises offsetting the cylinder axis from the crankshaft axis.

20. The method of claim 15, further comprising coupling a rolling element bearing between each piston and each cylinder.

21. The method of claim 15, further comprising installing a seal facing the oil and a seal facing the receiving cavity.

22. The method of claim 15, further comprising attaching a support band circumferentially around at least a portion of each piston.

Technical Field

The present disclosure relates to positive displacement pumps for moving liquids and slurries. More particularly, but not exclusively, the present disclosure relates to a diaphragm pump having an electric motor for actuating one or more diaphragms of the pump.

Background

Pumps may be used to facilitate the transfer of fluids including, but not limited to, liquids, slurries and mixtures. Thus, pumps (such as, for example, positive displacement pumps) may be designed to handle a range of fluid viscosities, including fluids containing relatively significant solids content, and to pump relatively harsh chemicals.

Positive displacement pumps may take many different forms, including, for example, positive displacement pumps that utilize a diaphragm or piston in connection with the intake and subsequent discharge of fluid from a chamber of the pump. For example, with respect to positive displacement pumps that are diaphragm pumps, such pumps typically include a pair of opposing diaphragms that reciprocate relative to each other along a common axis. Conventionally, these "dual diaphragm" pumps are pneumatically driven with high pressure air. Such a design may allow the pressure generated by the pump to be controlled by the pressure of the air in the system. Furthermore, such air operated diaphragm pumps are generally suitable for operation in potentially explosive environments, as pneumatic actuation generally prevents the generation of sparks.

However, Air Operated Diaphragm Pumps (AODP) have their disadvantages. For example, the high pressure air of the AODP is typically generated by an air compressor, which may be an add-on to the equipment required by the system and has an associated cost. In addition, reliance on pneumatic technology can result in poor net operating energy usage due to relatively significant losses in the generation, transport, and conversion of high pressure gases into mechanical work.

Thus, there remains an opportunity to create a pump that includes and improves upon the typical advantages of diaphragm pumps, while providing an alternative to relying on the inefficiency of pneumatically driven pumps.

Disclosure of Invention

This summary is provided to introduce a selection of concepts in a simplified form that are further described below in the detailed description. This summary is not intended to identify key features or essential features of the claimed subject matter, nor is it intended to be used to limit the scope of the claimed subject matter.

An aspect of an embodiment of the present disclosure is a diaphragm pump that may include a crankcase and a crankshaft positioned at least partially within the crankcase and rotatable about an axis of rotation. The diaphragm pump may include a piston coupled to a crankshaft by a connecting rod, the piston being reciprocally displaceable within a piston cylinder and along an axis of motion between a suction stroke and a discharge stroke, the axis of motion intersecting a connection between the piston and the connecting rod. The diaphragm housing may be coupled to an end of the piston cylinder and may be configured to at least partially define a pumping chamber and pump fluid through the pumping chamber as the piston reciprocates. The motion axis may not intersect the rotational axis of the crankshaft such that, relative to an arrangement in which the motion axis intersects the rotational axis, a peak value of piston-side load force encountered during a discharge stroke is reduced and a peak value of piston-side load force encountered during an intake stroke is increased to obtain a closer balance between the peak values of piston-side load force of the discharge stroke and intake stroke.

Another aspect of an embodiment of the present disclosure is a diaphragm pump system that may include a crankcase and a crankshaft positioned at least partially within the crankcase and coupled to an electric motor. Further, the crankshaft may be rotatable about an axis of rotation. At least three pistons may be radially disposed about the crankcase, each of the at least three pistons coupled to a throw of the crankshaft by a connecting rod. Additionally, each piston may be reciprocally displaceable within the piston cylinder and along an axis of motion between an intake stroke and a discharge stroke, the axis of motion of each of the at least three pistons intersecting the connection between the piston and the connecting rod. The diaphragm pump system may also include at least three diaphragm housings each coupled to an end of the piston cylinder and configured to at least partially define a pumping chamber and pump fluid through the pumping chamber as the piston reciprocates. Further, the axis of motion of each of the at least three pistons may not intersect the axis of rotation of the crankshaft such that the peak of piston-side load forces encountered during a discharge stroke is reduced and the peak of piston-side load forces encountered during an intake stroke is increased such that a closer balance is obtained between the piston-side load forces of the discharge stroke and the intake stroke relative to an arrangement in which the axis of motion intersects the axis of rotation.

Additionally, an aspect of an embodiment of the present disclosure is a diaphragm pump that may include a crankcase and a crankshaft positioned at least partially within the crankcase and rotatable about an axis of rotation. The diaphragm pump may include a piston coupled to a crankshaft by a connecting rod, the piston being reciprocally displaceable within a piston cylinder between a suction stroke and a discharge stroke. The diaphragm housing may be coupled to an end of the piston cylinder and may be configured to at least partially define a pumping chamber and pump fluid through the pumping chamber as the piston reciprocates. The piston cylinder may extend about a central longitudinal cylinder axis that intersects the axis of rotation. Additionally, the piston may be pivotally coupled to the connecting rod by a wrist pin positioned along a central longitudinal axis of the wrist pin that is parallel to the central longitudinal cylinder axis, linearly offset therefrom, such that a peak value of the piston-side load force encountered during the exhaust stroke is reduced and a peak value of the piston-side load force encountered during the intake stroke is increased relative to an arrangement in which the wrist pin is not linearly offset from the central longitudinal cylinder axis, so as to obtain a closer balance between the piston-side load forces of the exhaust stroke and the intake stroke.

These and other aspects of the disclosure will be better understood in view of the drawings and the following detailed description.

Drawings

The description herein makes reference to the accompanying drawings wherein like reference numerals refer to like parts throughout the several views.

Fig. 1 illustrates a diaphragm pump system according to an illustrated embodiment of the present disclosure.

Fig. 2 shows a perspective side view of a diaphragm pump according to an illustrated embodiment of the present disclosure.

Fig. 3 shows a cross-sectional view of the diaphragm pump taken along line 3-3 in fig. 2.

Fig. 4 shows a cross-sectional view of the diaphragm pump taken along line 4-4 in fig. 2.

FIG. 5 illustrates an exploded view of a diaphragm pump system and associated bracket according to an illustrated embodiment of the present disclosure.

FIG. 6 illustrates a side view of a diaphragm pump system and associated bracket according to an illustrated embodiment of the present disclosure.

FIG. 7 illustrates a side perspective view of a crankcase and piston member of a diaphragm pump according to an illustrated embodiment of the disclosure.

FIG. 8 shows a side view of a crankcase, an inner diaphragm casing, and certain piston members of a diaphragm pump according to an illustrative embodiment of the present disclosure.

Fig. 9 illustrates a graph showing outlet pressure at a common outlet of an electric diaphragm pump having three diaphragm housings as a function of crank angle according to an illustrated embodiment of the present disclosure.

Fig. 10 shows a graph showing the outlet pressure as a function of pump cycle in a prior art dual diaphragm pump.

FIG. 11A illustrates a cross-sectional view of a portion of an electric diaphragm pump having a linear offset slider-crank mechanism, according to an illustrated embodiment of the subject disclosure.

Fig. 11B shows an enlarged view of block 11B from fig. 11A depicting linearly offset centerlines of piston cylinders of an offset slider crank mechanism according to an illustrated embodiment of the subject disclosure.

FIG. 12 shows a graph depicting an example of the effect that an offset design of a slider crank mechanism can have on piston side loading as a function of crank angle.

FIG. 13 shows a graph depicting an example of the effect that an offset design of a slider crank mechanism can have on pump outlet pressure as a function of crank angle.

FIG. 14 shows a piston pin received in a piston pin cavity linearly offset from a corresponding cylinder axis.

Fig. 15A shows an enlarged view of a portion of a pump and an associated piston of a slider crank mechanism, the associated piston having an offset axis of motion, and the reciprocating displacement of the piston being guided by a linear guide.

Fig. 15B shows a front perspective view of a portion of a pump having a piston slidably coupled to a piston cylinder by a linear guide.

Fig. 16 shows an enlarged view of a portion of a diaphragm pump, wherein the axis of motion is angularly offset with respect to at least the axis of rotation.

The foregoing summary, as well as the following detailed description of certain embodiments of the present disclosure, will be better understood when read in conjunction with the appended drawings. For the purpose of illustrating the disclosure, certain embodiments are shown in the drawings. It should be understood, however, that the present disclosure is not limited to the arrangements and instrumentality shown in the attached drawings. Moreover, like reference symbols in the corresponding drawings indicate like or equivalent elements.

Detailed Description

Certain terminology is used in the foregoing description for convenience and is not intended to be limiting. Words such as "upper", "lower", "top", "bottom", "first" and "second" designate directions in the drawings to which reference is made. The terminology includes the words above specifically mentioned, derivatives thereof and words of similar import. In addition, the words "a" and "an" are defined to include one or more of the referenced item unless specifically mentioned otherwise. At least one of the phrases "following a listing of two or more items (such as" A, B or C ") means A, B or C, either alone or in any combination.

Fig. 1 illustrates a diaphragm pump system 50 according to an illustrated embodiment of the present disclosure. The diaphragm pump system 50 may include, among other components, a diaphragm pump 10, the diaphragm pump 10 being operatively coupled to a control system 12 and an actuator 14. While the embodiments discussed herein are discussed in terms of a diaphragm pump system including an electric diaphragm pump system, at least some features may also be applicable to various other types of pump systems, including but not limited to other types of pumps and positive displacement pumps, including but not limited to positive displacement pumps that utilize pistons rather than diaphragms to move fluid into/out of the pumping chambers of the pump. Additionally, at least some features of the diaphragm pump systems discussed herein may provide relatively significant advantages when compared to at least pneumatic diaphragm pump systems, including, but not limited to, increased energy efficiency in net operating energy usage.

According to some embodiments, the control system 12 may include, among other components, for example, an external embedded controller 11 communicatively coupled to a human machine interface 13. The external controller 11 may be configured to automate operation of the diaphragm pump 10 at least for dosing or dosing purposes. The external controller 11 may also be configured to add other cycle counting functions to the system 50. In addition, the external controller 11 may be configured to correlate the speed of the drive 14 (such as, for example, the motor speed) to the flow rate of the process fluid being pumped by the diaphragm pump. The external controller 11 may also include an override for a stall event for an extended period of time. Further, the control system 12 may be optional to supplement a motor drive, such as a Variable Frequency Drive (VFD) 15 configured to operate the drive 14.

As shown in at least fig. 1, the diaphragm pump 10 may be mechanically coupled to a driver 14. Although various types of drives 14 may be used, including, but not limited to, a variety of different types of engines and motors, according to the illustrated embodiment, the drives 14 are electric motors. Additionally, the drive 14 may be operatively coupled to the crankshaft 40 (fig. 4) of the diaphragm pump system 50 such that operation of the drive 14 may facilitate rotational displacement of at least the crankshaft 40 about a crankshaft axis (or "rotational axis") 100 (fig. 4). Further, as shown in at least fig. 1, according to certain embodiments, such operative coupling of the driver 14 to the crankshaft 40 may include a gearbox 16, and the gearbox 16 may be configured to adjust and/or control the relative speed and torque transmitted from the driver 14 to the crankshaft 40.

As shown in at least fig. 1-5, according to certain embodiments, the diaphragm pump 10 may include a crankcase 17, a plurality of diaphragm housings 18, a common inlet manifold 20 (fig. 5), a common outlet manifold 38, and a slider-crank mechanism 21 (fig. 3), among other components. Further, as shown by at least fig. 2, the crankcase 17 may include a lower crankcase 26 and an upper crankcase 28. As shown in at least fig. 4, the lower crankcase 26 may provide a lower crankcase volume 86. Additionally, a crankshaft 40 may protrude from the crankcase 17 for operative connection with the drive 14, as previously described.

While the number of diaphragm housings 18 may vary for different embodiments, the inventors of the subject disclosure have determined that an odd number of diaphragm housings (greater than one) may be preferred. Thus, the illustrated embodiment depicts, but is not limited to, a diaphragm pump 10 having three diaphragm assemblies 18. Further, each diaphragm casing 18 may be coupled to an adjacent piston 68 of the slider crank mechanism 21, as shown, for example, in fig. 3. In addition to the plurality of pistons 68, each reciprocally displaceable within a corresponding piston cylinder 60, the illustrated slider-crank mechanism 21 may also include a cam 82 (also referred to as a throw) of the crankshaft 40 and the connecting rod 62, as shown, for example, in fig. 4.

Additionally, in accordance with at least some embodiments, each of the diaphragm housings 18 may have generally similar components. Similarly, at least some of the components of the slider crank mechanism 21 associated with a particular diaphragm casing 18 may have the same configuration as other similar components of the slider crank mechanism 21 associated with another diaphragm casing 18. Thus, for example, each of the piston 68, the piston cylinder 60, and/or the connecting rod 62 of a slider crank mechanism 21 used with a particular diaphragm casing 18 may have similar configurations and features as similar components used with another diaphragm casing 18. Thus, it should be understood that, unless otherwise specified, parallel elements and associated features for these elements may be present for each of the diaphragm assemblies 18 and associated slider crank mechanisms 21, whether such parallel elements and features are actually visible in certain figures of the present disclosure or are explicitly discussed separately herein.

Each diaphragm housing 18 may include an outer housing 42 and an inner housing 44, the outer housing 42 may also be referred to as a fluid cap. As shown in at least fig. 3, at least an inner portion of outer housing 42 may generally define at least a portion of pumping chamber 46 of diaphragm housing 18. The pumping chamber 46 may be in fluid communication with the inlet 22 and the outlet 24 of the diaphragm housing 18. Thus, according to the illustrated embodiment, at least a portion of the process fluid entering the common inlet manifold 20 of the diaphragm pump 10 may enter the pumping chamber 46 of the diaphragm housing 18 through the inlet 22. Further, such process fluid may exit pumping chamber 46 through outlet 24 of diaphragm housing 18 and proceed to common outlet manifold 38 of diaphragm pump 10.

Additionally, as shown in FIG. 5, according to certain embodiments, a one-way check valve 48 may be functionally positioned near both the inlet 22 and the outlet 24 of each of the diaphragm housings 18. While various types of one-way check valves may be used, according to certain embodiments, the one-way check valve 48 is a ball valve. Additionally, according to certain embodiments, such ball valves may be gravity operated, and therefore do not include a biasing mechanism, such as, for example, a spring. However, alternatively, according to other embodiments, the one-way check valve 48 may include a biasing element such as, for example, a spring, in addition to other forms of biasing elements.

Fig. 3 shows a cross-sectional view taken along line 3-3 in fig. 2. Diaphragm housing 18 includes a diaphragm 80, diaphragm 80 being operable to vary the volume, and thus the pressure, within pumping chamber 46. Operation of diaphragm 80 may be used to draw process fluid into pumping chamber 46 through inlet 22, such as, for example, via displacing or flexing at least a portion of diaphragm 80 in a first direction to increase the volume within pumping chamber 46 and thereby decrease the pressure. Additionally, displacement or flexing of diaphragm 80 in a second, opposite direction may reduce the volume of pumping chamber 46 and thereby provide a pressure that may force at least a portion of the process fluid out of pumping chamber 46 through outlet 24.

Although various types of diaphragms may be used, according to certain embodiments, diaphragm 80 is a conventional flexible diaphragm. Additionally and optionally, according to certain embodiments, the septum 80 may be positioned in an opposite orientation between the inner housing 44 and the outer housing 42 as compared to the use of the septum in conventional AODP. According to certain embodiments, such as the embodiment shown in at least fig. 3 and 4, the diaphragm 80 may be positioned such that the arcuate shape of the annular flexible portion 83 of the diaphragm 80 is disposed in a direction generally away from the pumping chamber 46, and instead is directed in a general direction toward the receiving cavity 81 of the diaphragm housing 18.

The diaphragm 80 within the diaphragm housing 18 may be designed as a replaceable wear member. For example, in the illustrated embodiment, the diaphragm 80 is mechanically coupled to the second end 94 of the associated piston 68 via removable mechanical fasteners 74 (such as, for example, bolts). Further, according to certain embodiments, the mechanical fasteners 74 may extend through an inner washer 76 and an outer washer 78, the inner and outer washers 76, 78 being positioned on and supporting opposite sides of the diaphragm 80. For example, as shown in at least fig. 3, a radially inner portion of the diaphragm 80 may be secured between the inner and outer gaskets 76, 78. Inner gasket 76 and outer gasket 78 may be configured to provide stable and rigid support to at least adjacent portions of septum 80. In addition, a radially outward portion of the diaphragm 80 may fit securely between opposing sealing surfaces of the inner and outer housings 44, 42. Further, according to certain embodiments, the outer gasket 78 may be integrated into the diaphragm 80 such that the outer gasket 78 and the diaphragm 80 together have a unitary structure.

Further, as described below, the diaphragm housing 18 may be configured to minimize or avoid contamination of process fluid that may leak past the diaphragm 80, such as, for example, leaking past the diaphragm 80 due to damage or wear to the diaphragm 80. Such minimization or prevention of leakage past the diaphragm 80 may also minimize interruptions in the operation of the diaphragm 80 and, therefore, the diaphragm pump 10 and/or damage to the diaphragm 80 and, therefore, the diaphragm pump 10. Additionally, the diaphragm pump 10 can be similarly designed to minimize or avoid contamination of process fluid that may leak past the diaphragm 80.

More specifically, as can be seen in at least fig. 3, during a discharge stroke in which diaphragm 80 is forced axially away from rotational axis 100, process fluid may be pumped from pumping chamber 46 as the volume of pumping chamber 46 decreases. In the event that diaphragm 80 is damaged and/or diaphragm 80 fails, the pressure generated on the pumped fluid side of diaphragm 80 during a discharge stroke may tend to force at least a portion of the process fluid to flow past diaphragm 80 or behind diaphragm 80. However, in the illustrated embodiment, the receiving cavity 81 may be defined on the back side of the diaphragm 80. During normal operation, the receiving cavity 81 may include low pressure air, such as, for example, air at about ambient pressure (including, for example, no ambient air pressure of about 10 pounds per square inch (psi) as measured when the diaphragm pump 10 is not operating). The low pressure air may pass within the receiving cavity 81 of the separate diaphragm housing 18. Because each diaphragm 80 is in a different phase of its stroke at any one time, no significant pressure builds up in the receiving cavity 81.

Additionally, prior art diaphragm pumps typically use a high pressure working fluid, such as hydraulic fluid, stored behind the diaphragm to exert a fluid pressure on the back side of the diaphragm, which assists or fully drives the diaphragm. However, with such designs, leakage through the diaphragm can cause the working fluid to flow from the back of the diaphragm and into the process fluid, thereby contaminating the process fluid. However, unlike such designs, the receiving cavity 81 of the diaphragm housing 18 disclosed herein may only receive low pressure air because the diaphragm 80 is substantially fully mechanically actuated, such as, for example, by the corresponding piston 68 and components associated with the mechanical coupling of the piston 68 to the diaphragm 80. Thus, in accordance with certain embodiments of the subject disclosure, unlike prior designs that relied at least in part (if not entirely) on high pressure working fluid to drive the diaphragm, the annular flexible portion 83 of the diaphragm 80 is not driven by the working fluid, but instead may be substantially fully mechanically actuated.

The receiving cavity 81 may also be substantially sealed with respect to (from) a lubricant sump that may be within at least a portion of the crankcase 17 (such as, for example, lubricant within the crankcase cavity 86 that is used to reduce wear and distribute heat from the crankshaft 40 and connecting rod 62). For example, a seal assembly 72 (fig. 3) may bear against an outer surface of the piston 68. The seal assembly 72 may include, for example, one or more oil-facing seals and one or more containment cavity-facing seals, including but not limited to bellows seals and bi-directional seals. According to certain embodiments, the cavity-facing seal may be a bellows design (not shown) spanning between the second end 94 of the piston 68 and the piston cylinder 60. Seal assembly 72 may be configured and positioned to prevent lubricant from mixing with the process fluid even if the process fluid would leak past diaphragm 80 and reach receiving cavity 81.

Additionally, the containment chamber 81 may confine the process fluid during at least maintenance operations to minimize downtime of the diaphragm pump 10. For example, with simple removal of the outer housing 42 and the mechanical fasteners 74 of the diaphragm housing 18, as shown in at least fig. 4, the diaphragm 80 and the inner and outer washers 76, 78 may be removed, and the receiving cavity 81 may be easily and thoroughly cleaned.

With respect to operation of the slider crank mechanism 21, the piston 68 reciprocates along a piston axis that extends through the cylinder bore 59 of the piston cylinder 60 positioned between the crankcase 17 and the diaphragm housing 18. The piston 68 extends between a first end 92 and a second end 94 of the piston 68. The portion of the piston 68 proximate the crankcase 17 (i.e., the first end 92 of the piston 68) may include a piston pin cavity in which the piston pin 64 is positioned, the piston pin 64 attaching the piston 68 to the connecting rod 62.

The piston cylinder 60 may be removably mounted to the lower crankcase 26. As shown in at least fig. 3 and 4, according to certain embodiments, the piston cylinder 60 may be aligned with an aperture 88 of the lower crankcase 26 such that a portion of the piston cylinder 60 extends through the aperture 88 and toward the crankcase cavity 86. The piston cylinder 60 may also mate with the inner surface of the bore 88. Such an arrangement may provide increased stability to the piston cylinder 60 during operation of the pump 10. Additionally, such configurations may reduce the radial dimension of pump 10 via such positioning of piston cylinder 60, and as a result, piston 68, diaphragm 80, and outer housing 42 may be at a reduced radial position(s) from crankshaft 40. Additionally, as shown in at least fig. 8, the piston cylinder 60 may further include a shoulder 61, the shoulder 61 may be attached to a flat surface 138 of the crankcase 17, thereby providing increased stability to the piston cylinder 60 during operation of the pump 10 and improving ease of access and disassembly.

According to certain embodiments, the piston 68 and the piston cylinder 60 may be designed for controlled metal-to-metal sliding contact. In addition, one or both of the piston 68 and the piston cylinder 60 may be surface treated, such as with a diamond coating, to control wear of one or both of the piston 68 and the piston cylinder 60. In other embodiments, rolling contact may be provided between the piston 68 and the piston cylinder 60, such as, for example, via a rolling element bearing, which is a recirculating ball track extending against a track.

Additionally or alternatively, a sleeve or support band 70 (fig. 7) may be positioned circumferentially around a portion of the piston 6a, which may minimize or prevent metal-to-metal contact between adjacent portions of the piston 68 and the piston cylinder 60. The sleeve 70, which may be capable of being replaced as a wear part, may be made from a variety of materials including, for example, polymers, ceramics, or metals. Exemplary polymers that can provide suitable wear properties over the requisite pressure and velocity range of piston 68 can include Torlon, among other materials^®Polyester reinforced resins and bronze filled Polytetrafluoroethylene (PTFE).

For example, fig. 7 shows, among other features, a sleeve 70 attached to a first piston 68 and another second piston 68 prior to attaching the sleeve to the piston 68. With respect to the second piston 68, as seen, the outer surface of the piston 68 includes a sleeve recess 150 formed into the piston 68, the sleeve recess 150 being configured for seating of the sleeve onto the piston 68. As also seen, according to certain embodiments, the sleeve recess 150 may be a portion of the outer surface of the piston 68 having a dimension (such as, for example, a diameter) that is different from, such as, for example, smaller than, a corresponding dimension of other adjacent portions of the piston 68. Additionally, while the sleeve recess 150 may be located at multiple locations along the piston 68, as shown in fig. 7, according to certain embodiments, the sleeve recess 150 may be located at a location where, following attachment of the sleeve 70 to the piston 6b, the sleeve 70 will cover the wrist pin 64 that attaches the piston 68 to the associated connecting rod.

As previously mentioned, and as shown in at least FIG. 4, crankshaft 40 may rotate about an axis of rotation 100. Similarly, the cam 82, which is offset relative to the crankshaft 40, includes a central axis 102, and the central axis 102 may be parallel to the axis of rotation 100 and offset from the axis of rotation 100. According to some embodiments, crankshaft 40 may include a two-part shaft. Further, cam 82 may be integral with first portion 41 of crankshaft 40, while second portion 43 of crankshaft 40 may form seat 108. The seat 108 may be secured in the lower crankcase 26 by a first bearing set 110, and the crankshaft 40 may be secured in the upper crankcase 28 by a second bearing set 112. Additionally, the upper crankcase 28 may include a seal 114 that extends around a portion of the crankshaft 40.

As partially shown in FIG. 4, connecting rod 62 may extend from a connection with piston 68 as previously described to a connection with cam 82 of crankshaft 40. While the link 62 may be connected to the cam 82 in a number of different ways, according to the illustrated embodiment, the link 62 is connected to the cam 82 by a bearing ring or journal bearing 84. While bearing ring 84 may be coupled to connecting rods 62 in a variety of ways, as shown by at least fig. 4, according to the illustrated embodiment, bearing ring 84 may be positioned within an aperture in connecting rods 62. The bearing ring 84 may also be configured to facilitate sliding movement between the connecting rod 62 and the cam 82 of the crankshaft 40. Further, according to the illustrated embodiment, each bearing ring 84 is vertically displaceable relative to each other along the cam 82 and centered on a central axis 102 of the cam 82.

As shown in at least fig. 3 and 4, extending through each piston cylinder 60 is a respective central longitudinal cylinder axis l 16. Additionally, according to certain embodiments, each piston 68 shares its central axis with its corresponding cylinder axis 116. Further, according to certain embodiments, the piston pin 64 may also be positioned on the cylinder axis 116. Alternatively, according to other embodiments, the piston pin 64 may be linearly offset from the cylinder axis 116, which may provide an offset feature to the slider crank mechanism 21 that may improve the balance of piston-side load forces and stresses that may be encountered during the exhaust and intake strokes of the diaphragm housing 18, as described below.

As also partially shown in fig. 3 and 4, the diaphragm housing 18 may be similarly oriented about a cylinder axis 116 of the associated piston cylinder 60. Additionally, the bearing ring 84, the connecting rod 62, the piston cylinder 60, and the piston 68 may be centered on a horizontal plane that, along with similar horizontal planes for other diaphragm housings 18, may be vertically displaced along the cam 82.

Additionally, according to certain embodiments, each cylinder axis 116 for the diaphragm casing 18 is perpendicular to the rotational axis 100 of the crankshaft 40. Further, according to certain embodiments, the cylinder axes 116 of the diaphragm housings 18 may also be substantially equally radially spaced about the axis of rotation 100. For example, with respect to fig. 3, according to certain embodiments in which the diaphragm pump 10 includes three diaphragm housings 18, each cylinder axis 116 is disposed 120 degrees from the other cylinder axis 116. Because all three connecting rods 62 of the diaphragm housing 18 are disposed on the same cam 82 and equally spaced about the axis of rotation 100, the reciprocating motion of the respective pistons 68 are 120 degrees out of phase with each other. Thus, if the piston 68 of the first diaphragm housing 18 is at 0 degrees of its reciprocal cycle, the piston 68 of the second diaphragm housing 18 is at 120 degrees of its respective reciprocal cycle, and the piston 68 of the third diaphragm housing 18 is at 240 degrees of its respective reciprocal cycle. Similarly, for certain embodiments including five diaphragm casings, each piston may be disposed at approximately 72 degrees from its adjacent piston.

Fig. 5 illustrates an exploded view of an example diaphragm pump 10 and associated bracket 30 according to an illustrated embodiment of the present disclosure. As shown in the embodiment depicted in fig. 5, the diaphragm pump 10 may include the drive 14 and the gearbox 16 in a vertical orientation relative to the crankcase 17 and the bracket 30, with the drive shaft 19 of the drive 14 oriented to be directly or indirectly coaxially coupled with the crankshaft 40. Also shown in FIG. 5 is an exploded view of the diaphragm housing 18, as previously mentioned, the diaphragm housing 18 may each include at least an outer housing 42, an inner housing 44, a diaphragm 80, and a mechanical fastener 74. Also shown are common inlet manifold 20 and common outlet manifold 38, and one-way check valves 48 in operative communication with common inlet manifold 20 and common outlet manifold 38, respectively. Additionally, fig. 5 shows a three-leg bracket 30, wherein individual legs of the bracket 30 are disposed around the crankcase 17 at locations between adjacent diaphragm housings 18. Such legs of the bracket 30 may secure the pump 10 to a horizontal work surface with minimal work surface footprint.

Fig. 6 illustrates a side view of an alternative bracket 30' mounted diaphragm pump 10 according to at least one embodiment of the subject disclosure. The rack 30' depicted in fig. 6 is different from the rack 30 of fig. 5, and may include an upper rack portion 31, a lower rack portion 32, a rack base 34, and a plurality of supports 36. The membrane pump 10 may be attached to the bracket 30' at the upper part bracket portion 31 and/or the lower bracket portion 32. The bracket base 34 may be used to secure the diaphragm pump 10 to a work surface or floor, among other surfaces. Additionally, the rack base 34 may be configured for relatively easy pick up and movement by a forklift or other cart.

As indicated by at least fig. 5 and 6, the diaphragm pump 10 may be configured to be supported by the bracket 30, 30' in a substantially vertical orientation. Thus, the rotational axis 100 (fig. 5) of the crankshaft 40 and the drive shaft 19 of the drive 14 may also be disposed in a generally vertical direction. Moreover, such an orientation may accommodate a drive shaft 19 of the drive 14 that is substantially coaxial with the rotational axis 100 of the crankshaft 40. Such a vertical orientation of the diaphragm pump 10 may provide a number of advantages, including, for example, significantly reduced floor space for a work site and horizontal access to the pump 10 (which may be relatively free of other pump equipment), which may benefit the ability to perform maintenance on the pump 10, including replacement, repair, and/or cleaning of the pump 10 and/or components of the pump 10. Additionally, such a vertical orientation of the diaphragm pump 10 may allow the one-way check valve 48 to operate based on gravity, which may potentially reduce the number of components of the check valve 48, including, for example, avoiding a spring to bias a ball within the check valve 48. However, while the driver 14 depicted in fig. 1, 5, and 6 is shown mounted in a vertical orientation, the driver 14, as well as other components of the diaphragm pump system 50, may be mounted in a variety of other orientations.

FIG. 7 illustrates a side perspective view of the crankcase 17 and piston 68 of the diaphragm pump 10 according to an illustrated embodiment of the present disclosure. Furthermore, fig. 7 depicts at least the lower and upper crankcases 26, 28, wherein two of the pistons 68 protruding from the lower and upper crankcases 26, 28 can be seen.

As seen in fig. 7, according to the illustrated embodiment, the upper crankcase 28 may include a recessed section 130, and a plurality of first set of connector holes 132 for connecting portions of the upper crankcase 28 to the lower crankcase 26 at locations proximate to a curved surface 140 of the crankcase 17. The upper crankcase 28 may also include a second plurality of connector apertures 134 for connecting portions of the upper crankcase 28 to the lower crankcase 26 at locations proximate a planar surface 138 of the crankcase 17. The lower crankcase 26 may include a third set of connector apertures 136 for connecting the shoulder 61 of the piston cylinder 60 to an adjacent flat surface 138 of the crankcase 17. Additionally, the lower crankcase 26 may also include an outer wall 148, a flat surface 138, a curved surface 140, a first circulation port 142, and a second circulation port 144.

As seen in fig. 8, the connectors 160 may be positioned in at least a second set of connector apertures 134 (fig. 7), the second set of connector apertures 134 for connecting the upper crankcase 28 to the lower crankcase 26 at a location proximate the planar surface 138 of the crankcase 17. Additionally, a first circulation fitting 178 may be secured in the first circulation port 142 (fig. 7) and a second circulation fitting 180 may be secured in the second circulation port 144 (fig. 7).

Having described the structure of the diaphragm pump 10, operation will now be described further. In an exemplary embodiment, the driver 14 is an electric motor driven by an electric current, which may be controlled by the control system 12, for example. In response to receiving the electric current, driver 14 may facilitate rotation of drive shaft 19, drive shaft 19 being operatively connected to crankshaft 40 (with or without optional gearbox 16). Because of the offset between the axis of rotation 100 and the central axis 102 of the cam 82, rotation of the crankshaft 40 will generate reciprocating axial motion of each piston 68 along the cylinder bore 59 of its respective piston cylinder 60. As described above, by using a single cam 82 to drive each of the at least three pistons 68, in combination with the 120 degree spacing of the pistons 68 about the crankshaft axis 100 in this example, the motion of each piston 68 and the intake/exhaust cycle of each diaphragm 80 are 120 degrees or 240 degrees out of phase with the other pistons 68 and their associated diaphragms 80.

In certain embodiments, the electric diaphragm pump 10 is configured to provide a flow rate in the range of about 0 gallons per minute to about 300 gallons per minute at a pressure in the range of about 0 pounds per square inch (psi) to about 500psi through an inlet and an outlet having a diameter in the range of about 1/4 inches to about 6 inches. Embodiments of the present disclosure are also configured to provide dry lift of at least 15 feet. According to certain embodiments, the electric diaphragm pump is capable of performing wet lifts of at least about 20 feet and preferably at least about 30 feet.

Fig. 9 shows a graph showing the outlet pressure (dashed line) at the common outlet of an exemplary diaphragm pump 10 having three diaphragm housings 18 as a function of crank angle. As shown, the use of three diaphragms 80 with out-of-phase intake/exhaust cycles may generate a pressure profile that results in six outlet maximum pressure peaks per revolution of crankshaft 40 (P1-P6). As shown, the six maximum pressure peaks per 360 degree cycle of the diaphragm pump 10 are fairly horizontal, with the maximum pressure of these peaks only slightly different from the median pressure, as indicated by the solid line extending through the graph, and the minimum outlet pressure at the common outlet (M1-M4) also only slightly different from the median pressure, as shown.

Fig. 10 shows a graph showing the outlet pressure as a function of pump cycle in a prior art dual diaphragm pump. As shown in fig. 10, a prior art dual diaphragm pump may generate only two maximum pressure peaks per 360 degree cycle of the dual diaphragm pump. Further, the difference between the peak outlet pressure and the minimum outlet pressure per cycle with the prior art dual diaphragm pump is greater than the difference between the maximum outlet pressure and the minimum outlet pressure that can be achieved using the subject disclosed electric diaphragm pump 10 having three diaphragm housings 18.

A comparison of the pressure curves of fig. 9 and 10 shows a significant improvement in reduced pressure pulsations and increased mean pressure compared to conventional dual diaphragm pumps, which can be obtained with embodiments of the subject disclosed pump 10 that include three diaphragm housings 18. Furthermore, the subject disclosed three diaphragm pump 10 embodiments may reduce the amount of force on the system 50 by distributing the load across the three diaphragm assemblies 18 as compared to conventional two diaphragm designs.

Additionally, the diaphragm pump 10 may be designed to avoid pressure buildup when the diaphragm pump 10 is subject to a stall condition. Furthermore, diaphragm pumps are often used in industrial processes that require or otherwise cause temporary flow interruptions. Such flow disruption may be intentional, such as, for example, closing a valve to a nozzle via an operator, or may be unintentional, such as resulting from an accidental blockage in a flow path. In a typical air operated diaphragm pump, the air motor is designed such that the total flow interruption (commonly referred to as stall) avoids the build up of pressure in the process fluid (as air continues to be delivered to the pump).

With respect to the diaphragm pump system 50 of the subject disclosure, for example, the driver 14 (such as, for example, an electric motor) of the diaphragm pump 10 may be designed and controlled to decelerate, and even stop, when backpressure builds up during a stall event. For example, according to certain embodiments in which the drive 14 is an electric motor, the drive 14 may have a Pulse Width Modulation (PWM) based VFD controller 15 and may be capable of a constant torque mode, a constant speed mode, or a combination thereof. By programming the VFD controller 15 to operate at a desired or predetermined torque over a range of motor speeds, the drive 14 can be designed to vary its speed to maintain the desired torque, including running at a very slow speed. When faced with a stall event, the motor torque required by the drive 14 to drive the piston 68 typically increases as the discharge flow backs up to the outlet of the pump 14. Using a torque controlled drive 14 may facilitate a control system for the drive 14 to reduce the revolutions per minute (rpm) of the drive 14 so as not to exceed a predetermined threshold torque placed on the drive 14. By using this control, the rpm of the drive 14 can be reduced and actually stopped as long as the system places a torque on the drive 14 that exceeds a threshold. Thus, dangerously high back pressures in the discharge line from the diaphragm pump 10 can be avoided.

Additionally, according to certain embodiments, the driver 14 may be designed to maintain a constant speed up to a threshold torque. Thus, below the threshold torque, the drive 14 may be designed to maintain a selected speed even if the back pressure changes, which may otherwise affect the amount of torque on the drive 14. The constant speed of the driver 14 may be designed or selected to substantially maintain a selected flow rate of the diaphragm pump 10. Above the threshold torque, the drive 14 may be controlled to maintain the torque at the threshold by reducing the speed until the drive shaft 19 of the drive 14 rotates relatively very slowly, or stops in the event of a stall, in order to maintain pressure in the system, but not build up pressure.

In such embodiments, because the driver 14 is designed or configured to maintain pressure in the system 50 by maintaining torque at or below a selected threshold at the end of a stall event, the pressure of the pumped fluid is substantially immediately available when the stall condition is relieved, such as, for example, via opening a valve or flow in the discharge line. Further, the torque required by the drive 14 will fall below the selected torque threshold, the control system will actuate the increased rpm of the drive 14, and the discharge flow may progress from zero to the target flow rate. In other embodiments, the control system 12 may override and shut down the VFD controller 15 of the drive 14 if the stall event persists beyond a predetermined time limit, such as, for example, a one hour time limit.

Embodiments of the present disclosure may also exhibit relatively significant energy utilization efficiency. For example, with respect to electrical water efficiency (and more specifically, from the amount of electrical energy used to operate the driver 14 to the amount of kinetic energy transferred by the diaphragm pump 10 to the process fluid exiting the diaphragm pump 10), certain embodiments may achieve efficiencies greater than 50% over a large portion of the design operating range of the diaphragm pump 10. Further, according to certain embodiments, such efficiencies may be greater than 60%, and in some embodiments, efficiencies of about 65% may be obtained.

Embodiments of the present disclosure may also provide significantly reduced acoustic or noise profiles relative to those associated with many dual diaphragm pumps. Because the crankshaft 40 of the diaphragm pump 10 continuously rotates in one direction during operation (without a stall event), and the diaphragm 80 is coupled to the cam 82 by a substantially rigid connection, movement of the components of the pump 10, and in particular the movement of the diaphragm 80, is substantially smooth without intermittent sudden movements and accompanying acoustic shocks that are typically characteristic of operation of a dual diaphragm pump. Such designs of embodiments of the subject disclosure may also minimize or eliminate noisy dead motion connections and generated impulsive noise. Furthermore, the noise associated with the operation of the drive 14 (such as, for example, an electric motor) is generally quieter than the drive noise of an air motor from compressed air and AODP. Thus, the acoustic distribution of operation of embodiments of the present disclosure may provide significant advantages over conventional designs in terms of operation and working environment placement.

In addition, during operation, the degree of force acting on the diaphragm pump 10 during a suction stroke may be very different from these forces acting on the diaphragm pump 10 during a compression stroke. For example, at least some components of the diaphragm pump 10 used in displacement of the diaphragm 80 may experience relatively significantly higher levels of loading forces on the discharge stroke than those components encounter during the return/suction stroke. Accordingly, such components may experience higher wear rates over the exhaust portion of the stroke and require increased mechanical integrity for the exhaust portion of the stroke.

Referring to fig. 11A and 11B, according to certain embodiments, the slider-crank mechanism 221 may have one or more pistons 68 that are displaced in a reciprocating manner within the corresponding piston cylinder 60 along a motion axis 216, the motion axis 216 being offset from the rotational axis 100 of the crankshaft 40, and thus lying out of the plane thereof. According to some embodiments, the axis of motion 216 intersects the corresponding connection to the connecting rod 62 at the wrist pin 64 of the piston 68. Thus, in accordance with at least some embodiments, the axis of motion 216 extends through both the location where the center of the wrist pin 64 is located when the piston 68 completes the exhaust stroke and the location where the center of the wrist pin 64 is located when the piston 68 completes the intake stroke. Further, the location of the center of the piston pin 64 when the piston 68 completes the exhaust and intake strokes may be located on a central axis of the piston pin 68 that is generally located along the axis of motion 216 or shared by the axis of motion 216. According to certain embodiments, the degree of offset between the axis of motion 216 and the axis of rotation 100 of the crankshaft 40 may be at least the distance between the axis of motion 216 and the axis of rotation 100 of the crankshaft 40. Further, while fig. 11A and 11B depict the slider crank mechanism 221 as having three pistons 68 and three associated piston cylinders 60 and connecting rods 62, the number of pistons 68 and associated components used with the slider crank mechanism 221 may vary for different disclosures.

The offset of the axis of motion 216 relative to the axis of rotation 100 of the crankshaft 40 can be achieved in a number of different ways. For example, the slider-crank mechanism 221 depicted in fig. 11A and 11B is configured such that the axis of motion 216 along which the associated piston 68 is displaced in a reciprocating manner is linearly offset from the rotational axis 100 of the crankshaft 40. Such linear offset may be achieved, for example, by linearly adjusting the position of the movement axis 216 such that the movement axis 216 does not intersect the rotational axis 100 of the crankshaft 40 and is offset from the rotational axis 100. For example, and at least for purposes of discussion, the generally vertical orientation of the movement axis 216 associated with the third piston 68 shown in fig. 11B is offset in a generally horizontal direction (as indicated by direction "x" in fig. 11B) such that the movement axis 216 does not intersect the rotational axis 100 of the crankshaft 40, but is offset to the right of the rotational axis 100.

Such linear displacement of the axis of motion 216 of the slider crank mechanism 221 can be achieved in a number of different ways. For example, according to certain embodiments, cylinder bore 59 may be positioned or oriented such that central longitudinal axis 218 of cylinder bore 59 is linearly offset from rotational axis 100 of crankshaft 40. Since the axis of motion 216 associated with the reciprocating displacement of the piston 68 within the cylinder bore 59 may be coplanar with the central longitudinal axis 218 of the cylinder bore 59, an offset of the central longitudinal axis 218 relative to the axis of rotation 100 of the crankshaft 40 may result in a similar offset of the axis of motion 216 relative to the axis of rotation 100 of the crankshaft 40. Thus, according to such embodiments, the central longitudinal axis 218 and the corresponding axis of motion 216 of the cylinder bore 59 may be offset from the rotational axis 100 of the crankshaft 40 by substantially the same distance or magnitude, and in the same direction.

Alternatively, as previously described, and as shown in at least fig. 11A, the lower crankcase 26 may include one or more apertures 88 each sized and positioned to receive or otherwise couple to at least a portion of the piston cylinder 60. Such apertures 88 may be positioned and/or oriented such that a central longitudinal axis 217 of the apertures 88 is linearly offset from the rotational axis 100 of the crankshaft 40. Further, according to certain embodiments, such central longitudinal axis 217 of the bore 88 may be positioned such that when the piston cylinder 60 is attached to the lower crankcase 26 and the slider crank mechanism 221 is assembled, the motion axis 216 of the associated piston 68 is coplanar with the central longitudinal axis 217 of the bore 88, and the central longitudinal axis 217 of the bore 88 and the corresponding motion axis 216 are thus offset from the rotational axis 100 of the crankshaft 40 by substantially the same distance or magnitude.

As shown by at least fig. 11B, according to the illustrated embodiment in which the slider-crank mechanism 221 includes at least three pistons 68, the axis of motion 216 for each of the pistons 68 may be offset from the axis of rotation 100 of the crankshaft 40. Further, each axis of motion 216 may thus be oriented such that all three axes of motion 216 do not intersect at any common point.

Additionally, the amount of offset between the axis of motion 216 and the axis of rotation 100 of the crankshaft may be based on a variety of criteria, including, for example, but not limited to, stroke length. For example, according to certain embodiments, the axis of motion 216 may be offset from the rotational axis 100 of the crankshaft 40 by a distance of 0.1 inches to about 0.5 inches, and more specifically, by about 0.157 inches, among other distances.

The offset feature of the slider crank mechanism 221 may be configured to increase the duration of the discharge stroke during displacement of the piston 68 and associated operation of the diaphragm housing 118. Since the degree of forces and stresses encountered on the discharge stroke may generally be higher than those encountered on the intake stroke, increasing the amount of time spent on the discharge stroke may improve the balance between piston-side load forces and stresses that may be encountered during the discharge stroke and the intake stroke. Thus, the offset feature of the slider crank mechanism 221 may reduce the maximum forces and stresses experienced by the slider crank mechanism 221 and/or at least some components of the diaphragm casing 118. Such reduction in maximum force and stress may eliminate or reduce at least any need to over design the offset slider crank mechanism 221 and/or the diaphragm housing 118 of the pump 10, which may provide cost savings. Moreover, such improved balance of forces may facilitate better balancing expected wear on the diaphragm 80, as well as wear between at least the interface between the piston cylinder 60 and the associated piston 68, the sleeve or support band 70, and/or the associated linear guide assembly (fig. 15A and 15B), and thereby extend the useful life of such components.

For example, fig. 12 provides a graph depicting an example of piston side load as a function of crank angle of the slider crank mechanism 221 of the diaphragm pump 10, the diaphragm pump 10 having three levels of offset distance of the axis of motion 216 from the axis of rotation 100. With respect to a slider crank mechanism without an offset feature (e.g., "offset = 0 inches"), such as the slider crank 21 of fig. 3, the illustrated piston-side load force drops to about-80 pound-force (lbf) at its lowest during the intake stroke and reaches a maximum of about 600lbf during the exhaust stroke, as shown by the graph of fig. 12. In other words, in this example, without the offset feature, the maximum piston side load during the discharge stroke is about 7.5 times the maximum piston side load experienced during the intake stroke. However, for a slider crank 221 with an offset, when the axis of motion 216 is offset from the axis of rotation 100 by an offset distance of 0.2 inches in this example, an improved balance between piston side load forces between the intake stroke and the discharge stroke is shown, as indicated by a piston side load force on the intake stroke of up to about-130 lbf, and a maximum piston side load force during the discharge stroke of about 450 lbf. Thus, in this example, with a 0.2 inch offset between the axis of motion 216 and the axis of rotation 100, the maximum piston side load force during the discharge stroke drops to about 3.5 times the maximum piston side load force on the intake stroke. As further seen in this example, such balancing of piston-side load forces between the exhaust stroke and the intake stroke may be further enhanced by increasing the offset distance to 0.4 inches. Further, with an offset distance of 0.4 inches, the maximum piston side load forces for the intake stroke and the exhaust stroke in this example are about 200lbf and about 300lbf, respectively. Thus, with an offset distance of 0.4 inches, the maximum piston side load force of the discharge stroke drops to about 1.5 times the maximum piston side load force of the intake stroke. Thus, the variation in the offset distance may reduce the peak piston-side load force encountered during the discharge stroke while increasing the peak piston-side load force encountered during the intake stroke. Thus, a closer balance may be achieved between piston side load forces encountered during the discharge stroke and the intake stroke.

Thus, by providing a slider crank mechanism 221 with an offset feature, the diaphragm pump 10 can be designed and built using components that can withstand lower levels of force, as demonstrated by the example shown in FIG. 12. Further, referring to the data shown in FIG. 12, the diaphragm pump 10 may instead be constructed to withstand at least a maximum piston side load force of approximately 300lbf (as shown experienced by the example slider crank mechanism 221 with a 0.4 inch offset), rather than constructing the diaphragm pump 10 that may withstand at least a maximum piston side load force of approximately 600lbf (as shown experienced by the example slider crank mechanism 221 without the offset feature). Accordingly, such reduction in maximum force and maximum stress via incorporating the offset feature into the slider crank mechanism 221 may reduce, if not eliminate, any need to over design (such as, for example, oversize) at least the components of the slider crank mechanism 221, which may provide cost and size advantages in terms of components and manufacturing of the diaphragm pump.

The incorporation of the excursion feature into the slider crank mechanism 221, and the associated improved balancing of piston side load forces and stresses that may be encountered during the exhaust and intake strokes, may be provided without significantly changing the overall outlet pressure of the diaphragm pump 10. For example, fig. 13 provides a graph depicting an example of pump outlet pressure measured in pounds per square inch (psi) as a function of crank angle of the slider crank mechanism 21, 221 of the diaphragm pump 10, the diaphragm pump 10 having the same three levels of deflection as depicted in fig. 12. The outlet pressure shown in FIG. 13 may be a combined pressure effect of the diaphragm pump 10, the diaphragm pump 10 having three diaphragm housings 118 and thus three corresponding pistons 68. As shown in fig. 13, the total outlet pressure of the diaphragm pump 10 remains substantially the same for each of the three offset levels. Further, to the extent that fig. 12 and 13 show the maximum piston side load force and maximum/minimum pressure occurring at different crank angles, such differences may be attributed at least to the variation in the duration of the intake and discharge strokes, as previously described.

In addition, similar to fig. 9, fig. 13 also demonstrates that using an odd number of diaphragm housings 118 increases the number of pressure peaks that occur per operating cycle. Further, with respect to a diaphragm pump 10 having an odd number of diaphragm housings 118, the number of pressure peaks may equal twice the number of diaphragm housings 118. Thus, the data as depicted in FIG. 13 corresponds to an example diaphragm pump 10 having three diaphragm housings 118, and the number of pressure peaks occurring in each cycle is six, with three pressure peaks being generally about 115psi, and the other three pressure peaks being generally about 102 psi. In contrast, with a diaphragm pump having an even number of diaphragm housings, the number of pressure peaks is typically equal to the number of diaphragm housings, since each diaphragm has only one pressure peak. The additional pressure spike provided by using an odd number of diaphragm housings 118 may be a result of the increased duration of the overlapping time period in which multiple diaphragm housings 118 are subjected to a discharge stroke. Further, by increasing the duration of the discharge stroke of each diaphragm casing 118 through the use of the offset feature of the slider crank mechanism 221 of the subject disclosure, the duration of the discharge stroke experienced by multiple diaphragm casings 118 simultaneously may also be increased. Further, as previously described, the increase in the number of pressure peaks per cycle may enhance the loading shared by the diaphragm 80 of the pump 10 and increase the average pressure achievable by the pump 10.

While the foregoing example is discussed in terms of the linear offset of the axis of motion 216 of the slider crank mechanism 221 relative to the axis of rotation 100 of the crankshaft 40, the offset feature of the slider crank mechanism 221 may be provided in a variety of other ways. For example, according to certain embodiments, the piston pins 64 may be linearly offset from the corresponding cylinder axes 116, rather than offsetting the movement axes 216. For example, FIG. 14 shows a piston pin 64 received in a piston pin cavity 65 in a piston 68, the piston pin 64 attached to a connecting rod 62 coupled to a cam 82. As shown, cylinder axes 116 of the respective piston cylinders 60 (not shown) are positioned to intersect the rotational axis 100, wherein the rotational axis 100 is not positioned at the center of the cam 82, the cylinder axes 116 may also serve as axes of motion along which the pistons 68 are reciprocally displaced. However, the central longitudinal axis 67 of the piston pin 64 is positioned on the piston 68 at a location linearly offset from the cylinder axis 116, as indicated by distance "X" in fig. 14. According to the illustrated embodiment, the linear distance may be based on a distance from the central longitudinal axis 67 of the piston pin 64 and/or piston pin cavity 65 in a direction generally orthogonal to the cylinder axis 116. Moreover, such an offset of the piston pin 64 and/or piston pin cavity 65 may provide an adjusted angle of attack to the connecting rod 62 relative to the piston 68, which may at least increase the duration of the exhaust stroke, which again may facilitate improving the balance of forces experienced by the piston 68 during the intake and exhaust strokes.

Referring to fig. 16, according to other embodiments, instead of being linearly offset, the pump 10 may include a slider-crank mechanism 221 in which the axis of motion 216 of each diaphragm casing 18 is at least angularly offset relative to the axis of rotation 100 of the crankshaft 40 such that the axis of motion 216 does not intersect the axis of rotation 100. According to certain embodiments, such an offset of the axis of motion 216 may be achieved by angularly offsetting the central longitudinal axis 218 of the cylinder bore 59 of the piston cylinder 60 relative to at least the rotational axis 100 of the crankshaft 40. Such angular offset of the axis of motion 216 and the central longitudinal axis 218 of the cylinder bore 59 relative to at least the axis of rotation 100 may be achieved in a variety of ways. For example, according to certain embodiments, cylinder bore 59 may be formed in piston cylinder 60 such that central longitudinal axis 218 of cylinder bore 59 is angularly offset relative to central longitudinal axis 63 of piston cylinder 60. According to such embodiments, central longitudinal axis 63 of piston cylinder 60, rather than central longitudinal axis 218 of cylinder bore 59, may be positioned and oriented to intersect rotational axis 100. According to such embodiments, because the movement axis 216 may extend along the central longitudinal axis 218 of the cylinder bore 59, the movement axis 216 may also be offset relative to the rotational axis 100. Further, according to such embodiments, the piston pin 64 may be positioned along a central longitudinal axis 67 of the piston pin 64, the central longitudinal axis 67 being parallel to the movement axis 216, but linearly offset from the movement axis 216, as seen in fig. 16.

Alternatively, according to other embodiments in which central longitudinal axis 218 of cylinder bore 59 and, therefore, axis of motion 216 both extend along central longitudinal axis 63 of piston cylinder 60, piston cylinder 60 may be mounted to lower crankcase 26 via apertures 88 in a manner such that each of central longitudinal axis 63 of piston cylinder 60, central longitudinal axis 218 of cylinder bore 59, and axis of motion 216 is angularly offset from axis of rotation 100, without intersecting.

Fig. 15A shows an enlarged view of a portion of the pump 10 and the associated piston 68 of the slider crank mechanism 221, with the reciprocating displacement of the piston 68 guided by the linear guide or bearing assembly 202. According to the illustrated embodiment, the linear guide assembly 202 may include a bearing block 204, a plurality of balls or rollers (not shown), and a track 206. A plurality of balls or rollers, which may function as bearings, may be positioned between the bearing mount 204 and the track 206 such that the balls or rollers rotate as the bearing mount 204 linearly displaces along the track 206, thereby facilitating the linear displacement of the bearing mount 204 along the track 206. Further, the bearing seat 204 and the track 206 may have matching shapes to facilitate the bearing seat 204 remaining engaged with the track 206 and to at least assist in maintaining the plurality of balls or rollers in an operable position between the bearing seat 204 and the track 206.

As shown in fig. 15A and 15B, according to the illustrated embodiment, the rail 204 may be secured to an inner wall 208 of the piston cylinder 60, such as, for example, by one or more mechanical fasteners, including but not limited to one or more bolts. Further, according to certain embodiments, at least a portion of the track 206 may be recessed within a groove in the inner wall 208 of the piston cylinder 60. Similarly, the bearing block 204 may be fixed to the piston 68 such that the bearing block 204 linearly displaces with displacement of the piston 68. Thus, as the piston 68 is linearly displaced, such displacement of the piston 68 may be guided at least in a linear direction by linear motion of the bearing housing 204 along the rail 206. Further, according to certain embodiments, linear guide assembly 202 may provide a rolling interface between piston 68 and piston cylinder 60. Further, according to certain embodiments, at least a portion of the piston 68 may have a shape and/or size that may accommodate placement of at least a portion of the linear guide assembly 202 within the piston cylinder 60.

Additionally, similar to the embodiment discussed above with respect to fig. 14, fig. 15A also illustrates an embodiment in which the cylinder axis 216 of the corresponding piston cylinder 60 is positioned to intersect the rotational axis 100 of the crankshaft 40, wherein the rotational axis 100 is not positioned at the center of the cam 82, the cylinder axis 216 may also serve as an axis of motion along which the piston 68 is reciprocally displaced. However, similar to the embodiment discussed above with respect to fig. 14, the central longitudinal axis 67 of the piston pin 64 may be parallel to the axis of motion 116, but linearly offset therefrom, as indicated by the distance "X" in fig. 15A. Such deflection of the wrist pin 64 may also provide the connecting rod 62 with an adjusted angle of attack relative to the piston 68 that may at least increase the duration of the exhaust stroke, which may also facilitate improved balancing of piston-side load forces experienced during the intake and exhaust strokes.

While the linear guide assembly 202 is discussed above in relation to use with slider crank mechanisms 221 having offset features similar to at least those shown in fig. 14, the linear guide assembly 202 may also be used with other slider crank mechanisms that may have other types of offset features or configurations. Additionally, the linear guide assembly 202 may also be used with slider crank mechanisms that do not utilize an offset feature.

While the above examples are discussed with respect to a single piston cylinder and piston and their associated axes of motion, similar offset features may also be incorporated for any, if not all, of the other piston cylinders, pistons and associated axes of motion and/or associated diaphragm housings.

While the invention has been described in connection with what is presently considered to be the most practical and preferred embodiment, it is to be understood that the invention is not to be limited to the disclosed embodiment(s), but on the contrary, is intended to cover various modifications and equivalent arrangements included within the spirit and scope of the appended claims, which scope is to be accorded the broadest interpretation so as to encompass all such modifications and equivalent structures as is permitted under the law. Furthermore, it should be understood that while the use of the words "preferred," "preferably," or "preferred" in the description above indicate that the feature so described may be more desirable, it nonetheless may not be necessary and any embodiment lacking the same may be contemplated as within the scope of the invention, that scope being defined by the claims that follow. In reading the claims, it is intended that when words such as "a," "an," "at least one," and "at least a portion" are used, there is no intention to limit the claims to only one item unless specifically stated to the contrary in the claims. Moreover, when the language "at least a portion" and/or "a portion" is used, the item can include a portion and/or the entire item unless specifically stated to the contrary.