CN114151333A

CN114151333A - Screw compressor with split flow auxiliary compression and pulsation trap

Info

Publication number: CN114151333A
Application number: CN202110911749.1A
Authority: CN
Inventors: 黄秀保; 向·洋克斯
Original assignee: Haiba Blowing Machine Inc
Current assignee: Haiba Blowing Machine Inc
Priority date: 2020-09-08
Filing date: 2021-08-10
Publication date: 2022-03-08
Also published as: US20220074410A1

Abstract

A screw compressor with split flow assisted compression and pulsation traps for a split flow assisted compression and pulsation trap device (SECAT) assisted Internal Compression (IC) of a screw compressor reduces gas pulsations and induced vibrations and noise and improves operating efficiency under off-design conditions without the use of slide valves and/or series pulsation attenuators. The split flow assisted compression and pulsation trap includes an inner shell (e.g., an integral portion of the compression chamber) and an outer shell (e.g., a portion of the inner shell surrounding the compressor discharge) forming at least one diffusion chamber having a nozzle and a feedback region, at least one feedback flow loop being provided between the compressor chamber and the compressor discharge. The split flow assisted compression and pulsation trap automatically compensates for compression chamber pressure to meet the conditions of different outlet pressures (back pressures) to eliminate under-compression and/or over-compression when the discharge port is opened, partially recovers energy associated with under-compression (UC), isolates and attenuates gas pulsations and noise before the compression chamber opens to the discharge port.

Description

Screw compressor with split flow auxiliary compression and pulsation trap

Technical Field

The present invention relates generally to the field of rotary gas compressors, and more particularly to rotary screw compressors, commonly referred to as twin screw compressors, having dual intermeshing helical multi-vane rotors.

Background

Rotary screw compressors use two helical screws, called rotors, to compress gas. In a dry running screw compressor, a pair of positioning gears ensures that the male and female rotors each maintain precise position and clearance. In oil-immersed rotary screw compressors, an injected lubricating oil film fills the gaps between the rotors and the casing, both providing hydraulic sealing and transferring mechanical energy between the driving and driven rotors. Gas enters the compressor at the suction inlet and is trapped between the moving threads and the compressor housing, forming a series of moving cavities as the screw rotates. The volume of the moving cavity then gradually decreases causing the gas in the cavity to be compressed. The gas is discharged at the end of the screw compressor through a discharge port typically connected to a discharge muffler to complete the compression cycle. It is essentially a positive displacement compression mechanism but uses screw rotation rather than reciprocating motion, so the rate of volume change can be faster, fluid continuity better and the compressor size more compact than a conventional reciprocating piston type.

However, it has long been observed that screw compressors inherently produce gas (flow) pulsations with cavity passage frequency at discharge, and that the pulsation amplitude is particularly significant, whether under-compression (UC) or over-compression (OC), when operated at high pressure and/or under off-design conditions. As shown in fig. 1c, under-compression occurs when the gas pressure at the compressor outlet (discharge port) is greater than the gas pressure in the compressor cavity immediately before the discharge port opens. This can result in "explosive" backflow of gas from the outlet into the cavity, as shown in figure 1 a. On the other hand, when the pressure at the compressor outlet is less than the pressure in the compressor cavity before the discharge opening, as shown in FIG. 1d, over-compression occurs, resulting in an "explosive" forward flow, i.e., gas suddenly passes from the compression cavity into the compressor outlet as shown in FIG. 1 b. All fixed pressure ratio positive displacement compressors are subject to under-compression and/or over-compression because it is not possible to match only one fixed design pressure ratio to a varying system back pressure. Typical applications with variable pressure ratios include various refrigeration and heat pump systems and vacuum pumps. For example, as the ambient temperature rises or falls, the pressure ratios used in the refrigeration and heat pump systems must change accordingly. In general, the pressure ratio becomes much wider than 1, and the effect of OC and UC is further enhanced by the working fluid being a refrigerant causing the pressure increase required for operation. Another example of a condition requiring a wide range of pressure ratios is a vacuum pump, which is used to increase the vacuum in large vessels (e.g., to draw air from the vessel to the atmosphere), with the pressure ratios increasing with increasing vacuum in the vessel. For these applications, the energy losses and gas pulsations caused by UC and OC are significant, especially the latter, if the compressor outlet is not provided with a gas flow pulsation attenuator, which may damage downstream piping, equipment and cause severe vibration and noise in the compressor system.

To address the sequelae caused by the pressure ratio mismatch problem, gas flow pulsation dampers (also known as silencers), known in the industry as reactive (reactive) and/or absorptive (resistive), are typically installed at the discharge of the screw compressor, as shown in fig. 2a, to suppress and attenuate gas pulsations and induced vibrations and noise. Pulsation dampers are generally very effective in gas pulsation control, reducing pressure pulsations by 20-40dB, but are large in size and can cause other problems, such as more noise sources due to additional vibration surfaces, or sometimes fatigue failure damage to the damper structure, which can cause catastrophic damage to downstream components and equipment. At the same time, the exhaust dampers used today generate a large pressure (back pressure) loss, as shown in fig. 2b, resulting in a reduction of the overall efficiency of the compressor. For this reason, the screw compressor is often considered to have disadvantages of high gas pulsation, high vibration, high noise, low off-design efficiency, and large volume, as compared with a power type compressor such as a centrifugal compressor.

To overcome the problem of screw compressor pressure ratio mismatch from the source, the concept of a so-called slide valve has been widely explored since the 1960 s, as shown in figures 3a-3 b. There are cases where the slide valve concept appears in the title "device for regulating a screw rotary piston engine" (us patent No. 3,088,659) to h.r.nilsson et al, and in the title "screw rotary compressor without under-pressure and over-compression" (us patent No. 3,936,239) to n.shaw. The slide valve concept, commonly referred to as a variable volume ratio (Vi) scheme, uses a slide valve to mechanically change the internal volume ratio of the compressor, thereby changing the gas compression ratio of the compressor to meet the pressure requirements of different operating conditions and eliminating under-compression and/or over-compression, which are sources of exhaust gas flow pulsations and energy losses. However, slide valve systems are generally very complex in construction, costly, and have low reliability. Furthermore, they are not suitable for widely used dry screw applications because lubrication between sliding parts is essential.

Another technique that can also achieve the objectives of the slide valve variable volume ratio concept without its complexity and application limitations is the side branch Shunt Pulse Trap (SPT) technique disclosed by several of the present inventors, as shown in fig. 4a-4b (U.S. patent nos. 9,140,260, 9,155,292, 9,140,261, 9,243,557, 9,555,342 and 9,732,754). This technique uses a flowable gas to compensate for variable load conditions rather than moving solid mechanical parts that are sensitive to friction, fatigue failure, and response frequency. The SPT can achieve the same goal of a slide valve through an automatic feedback flow loop, i.e., communicating between the compressor cavity and the outlet (discharge port), compensating the pressure of the compression chamber by increasing or decreasing the gas in the cavity (as if the basketball were inflated or deflated to adjust its pressure) to eliminate under-or over-compression when the discharge port is open. The traditional SPT technology can effectively restrain low-frequency pressure pulsation in an under-pressure mode, and reduces energy consumption by eliminating the inherent back pressure loss of a series damper. However, it does not work well in the overpressure mode, especially for screw compressors operating over a wide range of pressure ratios.

It is therefore desirable to provide a new screw compressor design and construction that achieves low flow pulsations and low vibration noise at source and improves compressor off-design operating efficiency without the use of slide valves and external silencers at the exhaust, while having a small size, high reliability, and high efficiency operation in wide variable pressure ratio applications.

Disclosure of Invention

The invention aims to provide a screw compressor with a flow-dividing auxiliary compression and pulsation trap, which is a novel screw compressor design and structure, can realize low airflow pulsation and low vibration noise at the source and improve the efficiency of the compressor under the non-design working condition without using a slide valve and using an external silencer at an exhaust port, and has the advantages of small size, high reliability and high-efficiency operation in wide variable pressure ratio application.

The invention is realized by the following steps:

a screw compressor with split flow assisted compression and pulsation traps, comprising:

a compression chamber and a pair of meshed multi-spiral-vane rotors housed in the compression chamber, wherein the compression chamber has a suction port and a discharge port, wherein the rotors rotate in the compression chamber to form a series of moving chambers in the compression chamber for sucking and compressing gas and pushing the gas from the suction port toward the discharge port; and

a split flow assisted compression and pulsation trap (SECAT) apparatus comprising at least one diffusion chamber having a first stage gas nozzle providing fluid communication between a moving compression chamber inside the compression chamber and the diffusion chamber, and having a feedback region providing fluid communication between the diffusion chamber and a discharge outlet, wherein the split flow assisted compression and pulsation trap defines a first stage of a feedback flow loop,

the shunt-assisted compression and pulsation trap greatly reduces airflow pulsation, induced vibration and noise during operation, and improves the operation efficiency of the compressor under non-designed working conditions without using a series pulsation damper or/and a slide valve.

Wherein the first stage gas injection nozzles are located at least one rotor pitch from the suction port (i.e. completely isolated from the suction port) but before the exhaust port.

Wherein a second stage gas nozzle is included that is at least one rotor pitch distance from the first stage gas nozzle (i.e., completely isolated from the first stage gas nozzle), but is located before the exhaust port, thereby defining a second stage of a feedback flow loop.

Wherein a third stage gas nozzle is included that is at least one rotor pitch distance from the second stage gas nozzle (i.e., completely isolated from the second stage gas nozzle), but is located before the exhaust, thereby defining a third stage of a feedback flow loop.

The gas nozzles at all stages are circular in cross-sectional shape and have cross-sectional areas which are gradually transited along the axis of the nozzles and distributed in a converging or converging-diverging manner.

The gas nozzles at all levels are rectangular in cross section, and the cross section areas are gradually transited along the axis of the nozzle and are distributed in a convergent or convergent-divergent mode.

Wherein the convergent cross-section continuously transitions from a circular cross-sectional shape starting at the throat of the nozzle to a rectangular slot shape of the inner wall surface of the compression chamber of equal area, wherein the long sides of the rectangular slot-shaped nozzle at the inner wall surface of the compression chamber are parallel to the long sides of the slot shape of the moving compression chamber.

Wherein the convergent-divergent cross section continuously transitions from a circular cross-sectional shape starting at the throat of the nozzle to a rectangular slot shape of the inner wall surface of the compression chamber where the area gradually diverges, wherein the long sides of the rectangular slot-shaped nozzle at the inner wall surface of the compression chamber are parallel to the long sides of the slot shape of the moving compression chamber.

Wherein the stage gas nozzles are positioned at a distance away from the rotor axis and point in a tangential direction substantially the same as the direction of rotation of one of the rotors.

Wherein the pair of enmeshed multi-lobe rotors comprises a male rotor and a female rotor, with at least two gas injection nozzles per stage, at least one on the male rotor side and at least one on the female rotor side, and each pair of nozzles is positioned to open simultaneously to a corresponding moving male compression chamber and a moving female compression chamber in the compression chamber.

A screw compressor with split flow assisted compression and pulsation traps, comprising: a compression chamber and a pair of meshed multi-spiral-vane rotors housed in the compression chamber, wherein the compression chamber has a suction port and a discharge port, wherein the rotors rotate in the compression chamber to form a series of moving chambers in the compression chamber for sucking and compressing gas and pushing the gas from the suction port toward the discharge port; and

a split flow assisted compression and pulsation trap (SECAT) apparatus comprising at least one diffusion chamber having a first stage gas nozzle providing fluid communication between a moving compression chamber inside the compression chamber and the diffusion chamber, and having a feedback region providing fluid communication between the diffusion chamber and the surrounding atmosphere, wherein the split flow assisted compression and pulsation trap defines a first stage of a feedback flow loop,

the shunt auxiliary compression and pulsation trap can achieve deep vacuum of an air suction port during operation, greatly reduces air flow pulsation and induced vibration and noise, improves the operation efficiency of the compressor under the non-designed working condition, and does not need to use a series pulsation damper or/and a slide valve.

Generally, the present invention relates to a split assisted compression and pulsation trap (SECAT) for a screw compressor having a compression chamber with a suction port and a discharge port, and a pair of multi-lobe rotors housed in the compression chamber. The compression chamber rotors form a series of moving cavities therebetween for drawing, compressing and propelling gas in the compression chambers from the suction ports to the discharge ports. The split auxiliary compression and pulsation trap includes an inner housing that is an integral part of the compression chamber and an outer housing that is an outer enclosure. The inner casing adjacent the discharge port defines at least one diffusion chamber in which is housed at least one feedback flow circuit communicating between the moving compression chamber and the discharge port, each feedback flow circuit communicating with the moving compression chamber through at least one nozzle located at least one rotor pitch from the suction port (completely isolated from the suction port). Thus, the split auxiliary compression and pulsation trap can automatically compensate for the compression pressure in the compression cavity, as with the inflation or deflation of a basketball, the back pressure through the communication automatically adjusts the number of gas molecules added or reduced to the mobile compression cavity to meet the conditions of different outlet pressures (back pressures), thereby eliminating under-compression and/or over-compression when the exhaust port is opened, partially recovering the potential energy associated with under-compression (UC), and isolating and attenuating gas pulsation and noise before the exhaust port is opened.

The invention has the beneficial effects that: the design and structure of the novel screw compressor can realize low airflow pulsation and low vibration noise at the source and improve the efficiency of the non-design working condition of the compressor without using a slide valve and using an external silencer at an exhaust port, and simultaneously has the advantages of small size, high reliability and high-efficiency operation in wide variable pressure ratio application.

These and other aspects, features and advantages of the present invention will be understood with reference to the drawings and detailed description herein, and will be realized by means of the various elements and combinations particularly pointed out in the appended claims. It is to be understood that both the foregoing general description and the following brief description of the drawings, and the detailed description of the exemplary embodiments, are explanatory of exemplary embodiments of the invention, and are not restrictive of the invention.

Drawings

Fig. 1a and 1b are schematic diagrams of a triggering mechanism (instantaneous generation of airflow pulsation in the form of "compression wave-induced flow-expansion wave") of a conventional screw compressor for generating airflow pulsation when the compressor discharges air under-pressure and over-compression conditions.

FIGS. 1c and 1d are P-V plots of the associated energy losses for under-compression and over-compression conditions of a prior art screw compressor.

Figure 2a shows a schematic phase change diagram of a prior art screw compressor compression cycle with a series discharge muffler.

Figure 2b is a P-V plot of the compressor related energy loss of a prior art discharge tandem muffler (with back pressure).

Figures 3a and 3b show a typical design schematic of a prior art screw compressor with slide valve.

Fig. 4a shows a perspective view of a prior art side branch Shunt Pulsating Trap (SPT).

Fig. 4b is a cross-sectional view of a section (a-a) of the prior art shunt pulse trap of fig. 4 a. Figure 4b shows an alternative different shape of the nozzle.

FIG. 5 is a compression cycle phase change flow diagram with shunt assisted compression and pulse trap (SECAT) showing an under-compression condition and an over-compression condition in accordance with the present invention.

Fig. 6a is a cross-sectional view of a single stage split auxiliary compression and ripple trap according to a first exemplary embodiment of the present invention, showing an under-compression condition.

Fig. 6b is an expanded view of the cylinder inner wall of the single stage split assisted compression and pulsation trap of fig. 6 a.

Fig. 6c is a cross-sectional view of the single stage split assisted compression and pulsation trap of fig. 6a, showing an over-compression condition.

Figure 7a shows a side view and a top cross-sectional view of the nozzle with no shape transition from the nozzle throat to the inner wall of the compression chamber of the split flow assisted compression and pulsation trap and maintaining the same circular cross-sectional area.

Figure 7b shows a side view and a top cross-sectional view of the nozzle where the cross-sectional shape of the inner wall of the compression chamber from the nozzle throat to the split flow assisted compression and pulsation trap transitions from circular to rectangular but maintains the same cross-sectional area.

Figure 7c shows a side view and a top cross-sectional view of the nozzle where the compression chamber inner wall from the nozzle throat to the split flow assisted compression and pulsation trap not only has a cross-sectional shape transition from circular to rectangular, but also simultaneously increases the cross-sectional area (divergence).

Fig. 8a is a cross-sectional view of a two-stage split flow assisted compression and pulsation trap according to a second exemplary embodiment, showing both stages in an under-compressed state.

Fig. 8b is an expanded view of the cylinder inner wall of the two-stage split-flow assisted compression and pulsation trap of fig. 8 a.

Fig. 8c is a cross-sectional view of the two-stage shunt assisted compression and pulsation trap of fig. 8a, showing both stages in an over-compressed state.

Fig. 8d is a cross-sectional view of the two-stage shunt assisted compression and pulsation trap of fig. 8 a. Showing the first stage in an under-compression state and the second stage in an over-compression state.

Fig. 9a is a cross-sectional view of a single stage split auxiliary compression and pulsing trap according to a third exemplary embodiment, showing the split auxiliary compression and pulsing trap in a deep vacuum mode.

Fig. 9b is a cross-sectional view of a two-stage shunt assisted compression and pulsation trap according to a fourth exemplary embodiment, showing the shunt assisted compression and pulsation trap in a deep vacuum mode.

Detailed description of the preferred embodiments

Although specific embodiments of the present invention will now be described using reference to the accompanying drawings, it is to be understood that such embodiments are merely examples, and are merely illustrative of but a few of the specific embodiments that can represent the many possible applications of the principles of the present invention. Various changes and modifications obvious to one skilled in the art to which the invention pertains are deemed to lie within the spirit, scope and concept of the invention and are further defined by the appended claims.

It is also noted that while the dual rotor screw compressor for assisting in gas compression and attenuating gas pulsations is illustrated and described in the present invention, the principles may also be applied to a screw vacuum pump and/or other rotor combinations such as a single rotor screw compressor or a triple rotor screw compressor. The principle is also suitable for other gas media, such as refrigeration gas, and is also suitable for gas-liquid two-phase flow, such as an oil injection type screw compressor widely used for refrigeration. Furthermore, a screw expander is another variant for generating shaft power from a reduction in the medium pressure.

To illustrate the principles of the present invention, fig. 5 is a flow chart of a screw compression cycle of a "split assisted compression and pulsation trap" (secat) according to an exemplary embodiment of the present invention, showing the communication of the internal compression phase and the discharge port. Broadly speaking, split-flow assisted compression and pulsation traps are used to assist Internal Compression (IC), while isolating and attenuating flow pulsations and noise and increasing the efficiency of operation in off-design conditions without the use of slide valves and/or conventional series pulsation resistive mufflers. As shown in FIG. 5, the split flow assisted compression and pulsation trap involves a modification to the standard (conventional) screw compression cycle, from a series mode, i.e., from the series internal compression and under, over compression and pulsation attenuation shown in the prior art of FIG. 2a, to a parallel operation of internal compression and under, over compression and pulsation attenuation, such that both operate simultaneously and in concert over a longer time interval. Due to undervoltage Δ P_UC(＝P_outlet-P_cavity) Or an overpressure Δ P_OC(＝P_cavity–P_outlet) Any pressure differential that results in a compressor chamber pressure and a target outlet pressure will automatically trigger a feedback flow, i.e., a Δ P Induced Fluid Flow (IFF) between the compression chamber and the exhaust port, in a manner that increases or decreases the number of additional gas molecules entering or exiting the compression chamber, with the goal of minimizing the pressure differential (Δ P) between the compression chamber and the exhaust port before the discharge valve opens. This screw compression chamber pressure compensation is similar in principle to the regulation of the pressure in a fixed volume cavity by injecting or releasing gas into the cavity, such as the inflation or deflation of a basketball for sports activities. With a compound compression scheme with internal and split flow assisted compression and pulsation traps, any under or over compression at the compressor discharge opening will be compensated to a minimum, thus eliminating the need for downstream resistive mufflers (although appropriate resistive mufflers may be used if desired for broadband noise attenuation, such as gas venting to atmosphere for vacuum applications).

Referring to fig. 6a to 6c, a typical arrangement of a screw compressor 10 with a split assisted compression and pulsation trap (secat) device 50 according to a first exemplary embodiment is shown. Generally, the screw compressor 10 has two rotors 12 respectively integrated with two rotor shafts 11, wherein one rotor shaft 11 is driven by an external rotational driving mechanism (not shown). The rotors 12 are typically driven by a set of positioning gears (in the case of dry operation) or they are directly driven by each other (in the case of oiling). The rotor 12 is typically a pair of multi-lobed rotors, a male and a female, housed in a compression chamber 32, forming a series of moving chambers, e.g. 38 and 39, for drawing, compressing and propelling the drawn gas from a suction port 36 to a discharge port 37 of the compressor 10. The screw compressor 10 also has an inner shell 20 as an integral part of the compression chamber 32, with the rotor shaft 11 mounted in an internal bearing support structure (not shown). The housing structure further includes an outer housing 28 surrounding a portion of the inner housing 20 adjacent the discharge opening 37 to form at least one diffusion chamber 55.

As a novel and unique feature of the present invention, the split flow assisted compression and pulsation trap device 50 includes at least one flow nozzle (trap inlet) 51 branching from the compression chamber 32 to at least one diffusion chamber 55 and a feedback region (trap outlet) 58 in communication with the compressor outlet 37. In fig. 6b, the starting line of the flow nozzle (trap inlet) 51 is located at least one blade span (or pitch t) from the closing line of the suction inlet 36 at which the

compression chamber

38 or 39 moves and is positioned (distance d in fig. 6 b) as far away as possible from the axis of rotation 11 and pointing in the same general direction as the direction of the rotating rotor 12 to assist the rotor rotation (e.g. positioned parallel to the tangent of the direction of rotation). Fig. 6b also shows two types of nozzles 51 used: the left side is a nozzle with a rectangular cross-sectional shape, the cross-sectional area distribution of which is reduced along the axis; to the right are two nozzles having a circular cross-sectional shape with a decreasing cross-sectional area distribution along the axis. Fig. 6a shows the flow pattern of the under-compressed state, where large directional arrow 30 shows the direction of movement of the compression cavity that is pushed by the cavity formed by rotor 12 from suction port 36 to discharge port 37 of compressor 10, while the feedback flow IFF, shown by small directional arrow 53, passes from feedback region (trap outlet) 58 through diffusion chamber 55, then converges to flow nozzle (trap inlet) 51 and discharges into compression cavity 39, which is open to flow nozzle 51. On the other hand, fig. 6c shows a flow pattern over compression, where the large directional arrow 30 still shows the direction of movement of the compression cavity formed by the rotor 12, pushed from the suction port 36 to the discharge port 37 of the compressor 10, while the feedback flow IFF, shown by the small arrow 54, passes from the compression chamber 39, now open to the discharge port, through the flow nozzle 51 to the diffusion chamber 55 and is released to the trap outlet 58 where it meets the discharge flow 30.

When the screw compressor 10 equipped with the split flow assisted compression and pulsation trap apparatus 50 of the present invention is operated in under-pressure and/or over-compression conditions, not only are the flow pulsations and induced noise transmitted from the screw compressor exit to the downstream flow reduced, but the internal flow field (and hence the adiabatic efficiency under off-design conditions) is also improved. The principle of operation of the split flow assisted compression and pulsation trap apparatus 50 of the present invention can be described as follows. In the under-compression mode, as shown in FIG. 6a, the split-flow assisted compression and pulsation traps are typically operated from the gas pressure in the compression chamber 39 to the minimum of the applicationWorking pressure P₁(but well below the maximum pressure) begins to assist in internal compression at that point. When the air pressure is P₁The "moving compression chamber" 39 suddenly has a pressure P₄The shunt assisted compression and pulsation trap's trap inlet 51 opens, triggering a shock tube snap-open like reaction (as disclosed in commonly owned U.S. patent No. 9,155,292). This produces a transient gas pulsation in the form of a "compressional wave-induced flow-expansion wave" at the abruptly open nozzle throat 51, with the compressional wave C W (not shown) and induced flow IFF 53 passing into the compression chamber 39, and the expansion wave EW (not shown) propagating outwardly from the nozzle 51 toward the trap outlet 58 and compressor discharge 37.

The split flow auxiliary compression and pulsation traps have several advantages over screw compressors with series connected conventional silencers. First, the use of the nozzle 51 more efficiently delivers the required mass to the under-pressure compression chamber 39 in a "starved" condition, reducing the time to fill the compression chamber 39 and the occurrence of gas flow pulsations upon discharge. As shown in fig. 6a and 5, the required gas mass flow 53 is first "borrowed" from outlet region 37 and then "returned" to outlet region 37 by a split feedback flow loop, so that induced flow 53 is not lost in the process. The flow rate of feedback stream 53 is designed to accomplish compensatory internal compression prior to discharge, such that pressure differential Δ P occurs at discharge_UCOr Δ P_OCMost of it is eliminated or reduced to near zero as shown in fig. 5. The jet speed of the nozzle throat can be close to or equal to the local gas sound speed (when delta P)_UCLarger) than the moving chamber 39, so this solution can be used in high speed dry screw compressors where the slide valve variable volume ratio design does not work properly. Second, from a noise reduction perspective, as long as the nozzle throat 51 is blocked (reaches sonic velocity), the nozzle 51 acts as a trap to isolate the high velocity jet noise within the compression cavity 39 prior to discharge so that no compression waves CW and jet-induced noise can escape or propagate upstream through the nozzle throat 51. When the nozzle throat 51 is unobstructed (throat gas velocity less than sonic), the propagation of compression wave CW and jet noise within the chamber 39 is also greatly reduced because of the noiseThe throat area through which the escape is very small. Furthermore, the gas velocity at the diffuser end of the nozzle 51 leading to the diffuser chamber 55 and downstream outlet 37 is much lower than the throat velocity, and therefore the flow induced noise is also much lower. Third, from an energy conservation perspective, the energy loss traditionally associated with under-compression, as shown by the shaded area in fig. 1c, can now be partially recovered (positive work) because the torque generated by the high velocity jets 53 is now used to propel the rotor 12 in the same direction of motion, as shown in fig. 6b, as a Pelton Wheel (Pelton Wheel) for paddlewheel work generation. In the conventional tandem scheme shown in prior art fig. 1a and 2a, the backflow jets generally act in opposite directions of rotor rotation, resulting in negative work being done on the compressor system.

On the other hand, the basic operating principle of the shunt assisted compression and pulsation trap device 50 for over-compression mode is different. As shown in fig. 6c and 6b, since the split-flow assisted compression and pulsation trap is designed to assist the gas pressure P from the compression chamber 39₁Slightly exceeding the minimum applied working pressure P at the outlet 37 of the compressor 10₂At the moment of the compression, the auxiliary internal compression is started, since the gas pressure is P₁The "moving compression chamber" 39 opens abruptly to the trap inlet 51 of the split-flow assisted compression and pulsation trap, at a pressure P₂Just slightly below P₁Is a small perturbation Δ P_OC(＝P₁–P₂) And thus a shock-like reaction does not occur. Instead, only a small feedback flow IFF 54 is generated (as indicated by the small directional arrow in fig. 6 c): from the compression chamber 39 to the nozzle 51, through the diffusion chamber 55 and to the trap outlet 58 where it meets the discharge. Because the volume of the compression cavity 39 is gradually reduced, i.e., the internal compression is gradual in nature (rather than abrupt), the magnitude of the induced flow IFF 54 is much smaller in the case of over-compression than the induced flow IFF 53 produced when under-compression; thus, the injection noise caused by the flow 54 in the event of over-compression is much lower.

To optimize the

feedback flow

53 or 54 in either direction in the nozzle 51 between the compression chamber 39 and the diffusion chamber 55, a minimum of one nozzle may be used on the male rotor side or/and the female rotor side to compensate for the pressure in the compression chamber 39, the nozzle may be selected in the form of a plurality of circular holes and/or slots arranged parallel to the thread sealing line of the chamber 39 or a transition from circular to slotted. Both are shown in fig. 6b for illustrative purposes. Furthermore, if a circular cross-sectional shape nozzle is used, the cross-section of the throat 59 into the compression chamber 39 may be designed to be circular (fig. 7a, no shape and area change) of the same cross-sectional area as the nozzle throat 59, or gradually transition from circular to trough (fig. 7b, with shape change but no area change), with the trough direction parallel to the thread seal line of the chamber 39, or with a gradually increasing cross-sectional area (fig. 7c, both shape and area change), forming a so-called Laval nozzle. Replacing the circular cross-sectional shape (fig. 7a) with transition slots as shown in fig. 7b and 7c reduces the stage spacing of the multi-stage split auxiliary compression and pulsation traps (shown in fig. 6 b) defined perpendicular to the rotor seal line, thereby allowing more time for the second stage split auxiliary compression and pulsation trap operation. Furthermore, the elongated shape of the nozzle slot on the inner wall surface of the cylinder will facilitate flow exchange with the elongated shaped compression chamber 39, i.e., increase the flow efficiency of the feedback flow 53 for under-compression or the feedback flow 54 for over-compression, especially for high speed dry screw applications where the exchange time is short.

The first stage split auxiliary compression and pulsation traps are sufficient to cover the entire composite compression phase if the pressure ratio and its variation range (the range of under-compression and over-compression) are relatively small, i.e. when the distance between the nozzle 51 opening to the discharge 37 is less than one blade span or pitch t (as shown in fig. 6 b). However, for some applications where the pressure ratio and its range of variation are large, a two-stage split flow assisted compression and pulsation trap may be used to cover the compound compression stage when the distance between the first nozzle opening closing and the discharge opening is greater than one blade span or pitch t. The principle is that each compression cavity is always communicated with the exhaust port of the compressor at any moment of the composite compression phase, but all the compression cavities are never communicated with each other. Based on this principle, the start of the second stage nozzle should be located at least one pitch t from the end of the first stage nozzle and within the last pitch before the exhaust port opens. Also, if the two-stage split auxiliary compression and pulsation traps are not sufficient to cover the entire composite compression phase, three-stage split auxiliary compression and pulsation traps may be used.

Referring to fig. 8a to 8c, a second exemplary embodiment of a screw compressor 10 according to the present invention with a split assisted compression and pulsation trap (secat) device 60 is shown: two-stage shunting assists in the typical arrangement of compression and pulsation traps. The construction of the screw compressor 10 and the first stage of the split assisted compression and pulsation trap device 60 is the same as the split assisted compression and pulsation trap device 50 described above. However, a second stage split assisted compression and pulsation trap arrangement 60 is added, which further comprises at least one further side branch circuit: from the compression chamber 32 through the nozzle 61 (trap inlet) to at least one diffusion chamber 63 and connected to a feedback zone (trap outlet) 68 and communicating with the compressor outlet 37. As shown in fig. 8b, the first stage nozzle 51 (trap inlet) is still at least one lobe (one pitch t) from the line of closure of the suction inlet 36, while the start of the second stage nozzle 61 is at least one pitch t from the line of closure of the first stage nozzle 51, both stages being as far away from the axis of rotation 11 as possible (distance d in fig. 8 b) and pointing in the same direction as the rotating rotor 12 to assist its rotation. Fig. 8a shows the flow conditions in which both stages are in an under-compression mode, wherein the large directional arrows 30 show the direction of movement of the compression cavity formed by the rotor 12 from the suction port 36 to the discharge port 37 of the compressor 10, while the feedback flows 53 and 63, shown by the small directional arrows, pass from the feedback zone (trap outlet) 58 through the

diffusion chambers

55 and 65, then converge to the

flow nozzles

51 and 61 and discharge into the

compression chambers

38 and 39, respectively. On the other hand, fig. 8c shows the flow conditions in both stages in an over-compression mode, where the large directional arrows 30 still show the direction of movement of the compression cavity formed by the rotor 12 from the suction port 36 to the discharge port 37, and the feedback flows 54 and 64 are shown as small directional arrows: from the

compression chambers

38 and 39, which were just opened to

nozzles

51 and 61, through

diffusion chambers

55 and 65 and finally to trap

outlets

58 and 68, merging with the exhaust stream 30. Fig. 8d shows the flow conditions with the first stage under-compressed and the second stage over-compressed, wherein the feedback flow 54 of the first stage passes from the feedback zone (trap outlet) 58, through the diffusion chamber 55, then to the flow nozzle 51, and finally into the cavity 38, while the feedback flow 64 of the second stage passes from the just-opened compression cavity 39 to the nozzle 61, through the diffusion chamber 65, and into the trap outlet 68, and finally joins the exhaust flow 30.

In addition to the applications discussed above for the first and second exemplary embodiments for a two-port (only inlet and exhaust) configured screw compressor, a three-port configured screw vacuum pump may be used for deep vacuum pumping applications. In the vacuum pump embodiment, the suction port of the compressor is connected to the process or vessel in which the deep vacuum is to be generated, while the discharge port of the compressor is connected to the atmosphere through a muffler. In addition, a third port is added which is also open to atmosphere and allows cool air to enter the compressor cavity through a split auxiliary compression and pulsation trap to extend the pressure ratio range, for example, from about 4/1 maximum dry screw compressor pressure ratio to about 20/1 or more.

Referring to fig. 9a and 9b, a screw compressor 10 with split assisted compression and pulsation trap (secat)

devices

70 and 80 according to a third and fourth exemplary embodiment is shown, respectively: single and double stage shunting aid in typical arrangements of compression and pulsation traps. The configuration of the screw compressor 10 with the split assisted compression and

pulsation trap apparatus

70 and 80 differs from that of the split assisted compression and pulsation trap apparatus 50 and 60 (of the first and second embodiments) in that: the addition of a third port 77 (instead of the feedback region 58 in fig. 6 a) connects the compressor chambers 38 and/or 39 directly to the atmosphere 78 by shunting the auxiliary compression and

pulsation trap devices

70 and 80, rather than merging with the compressor outlet 37. Fig. 9a shows a typical mode of operation of a single stage split flow assisted compression and pulsation trap 70. As shown in fig. 9a, when the operating pressure ratio is less than the design pressure ratio of compressor 10, flow (not shown) is first released from compression chambers 39 through nozzles 51, then through diffusion chamber 55 to port 77 and into atmosphere 78, prematurely breaking out of the over-compression to conserve energy and reduce exhaust gas flow pulsations. And when the operating pressure ratio is greater than the design pressure ratio of compressor 10, the flow direction (not shown) is automatically switched to draw cooler atmospheric air from port 77 through diffusion chamber 55 and nozzle 51 into compressor cavity 39. The mixing of the ambient air with the inner compressed warmer cavity air will cause the compressor to reach a maximum pressure ratio much higher than its normal operating range, for example from about 4/1 to about 20/1 or higher.

Accordingly, various embodiments of the present invention provide advantages over the prior art. For example, a screw compressor with a split flow assisted compression and pulsation trap (SECAT) in parallel with compression within the compressor helps to eliminate under-compression and/or over-compression (a source of discharge pulsations and energy losses) when the discharge port is open. Screw compressors with split assisted compression and pulsation traps (SECAT) can be as effective as slide valve variable volume ratio designs, but without mechanically moving parts, nor the requirement for oil injection applications, can be used for high speed dry screw compressor applications. Screw compressors with split flow assisted compression and pulsation traps (SECAT) can be an integral part of the compressor shell, thus eliminating the pulsation silencers connected in series at discharge to make it compact in size. Screw compressors with split flow assisted compression and pulsation traps (SECAT) can achieve energy savings over a wide range of pressure ratio variations. Screw compressors with split flow assisted compression and pulsation traps (SECAT) are capable of reducing air flow pulsations and induced vibratory noise over a wide range of pressure ratio variations. Screw compressors with split flow assisted compression and pulsation traps (SECAT) enable energy savings and gas pulsation attenuation over a wide range of speeds and cavity pass frequencies. Screw compressors with split assisted compression and pulsation traps (SECAT) are capable of achieving the same level of adiabatic off-design efficiency as slide valve technology over a wide range of pressure ratio variations and speeds.

It is to be understood that this invention is not limited to the particular devices, methods, conditions or parameters of the example embodiments described and/or illustrated herein, and that the terminology used herein is for the purpose of describing particular embodiments only and is intended to be exemplary. Accordingly, the terms are intended to be broadly construed and are not intended to unnecessarily limit the claimed invention. For example, as used in the specification, including the appended claims, the singular forms "a," "an," and "the" include plural, and the term "or" means "and/or" and refers to particular digits. Unless the context clearly dictates otherwise, a numerical value includes at least that particular numerical value. Additionally, any methods described herein are not intended to be limited to the order of the steps described, but may be performed in other orders, unless otherwise specifically indicated herein.

While the claimed invention has been shown and described in detail in the foregoing for the purpose of illustration, it will be apparent to those skilled in the art that various modifications, additions and variations can be made thereto without departing from the spirit and scope of the invention as defined in the following claims.

Claims

1. A screw compressor with split flow assisted compression and pulsation traps, comprising:

2. The screw compressor of claim 1, wherein the first stage gas injection nozzle is located at least one rotor pitch from suction (i.e., completely isolated from suction), but before the discharge.

3. The screw compressor of claim 1, further comprising a second stage gas nozzle spaced at least one rotor pitch from the first stage gas nozzle (i.e., completely isolated from the first stage gas nozzle), but before the discharge port, thereby defining a second stage of a feedback flow loop.

4. The screw compressor of claim 1, further comprising a third stage gas nozzle spaced at least one rotor pitch from the second stage gas nozzle (i.e., completely isolated from the second stage gas nozzle), but before the discharge port, thereby defining a third stage of a feedback flow loop.

5. The screw compressor according to claim 1, wherein each stage of gas nozzles is circular in cross-sectional shape with a converging or converging-diverging distribution of cross-sectional area that gradually transitions along the nozzle axis.

6. The screw compressor according to claim 1, wherein each stage of the gas nozzle has a rectangular cross-sectional shape with a converging or converging-diverging distribution of cross-sectional areas that gradually transition along the nozzle axis.

7. The screw compressor according to claim 5, wherein the convergent cross-section continuously transitions from a circular cross-sectional shape starting at the throat of the nozzle to a rectangular slot shape of an inner wall surface of the compression chamber of equal area, wherein the long sides of the rectangular slot-shaped nozzle at the inner wall surface of the compression chamber are parallel to the long sides of the slot shape of the moving compression chamber.

8. The screw compressor according to claim 5, wherein the convergent-divergent section continuously transitions from a circular sectional shape starting at the throat of the nozzle to a rectangular slot shape of an inner wall surface of the compression chamber having a gradually diverging area, wherein the long side of the rectangular slot-shaped nozzle located at the inner wall surface of the compression chamber is parallel to the long side of the slot shape of the moving compression chamber.

9. The screw compressor of claim 1, wherein the stage gas nozzles are positioned a distance away from the rotor axis and point in a tangential direction substantially the same as the direction of rotation of one of the rotors.

10. The screw compressor according to claim 1 wherein the pair of enmeshed multi-lobe rotors comprises a male rotor and a female rotor, with at least two gas injection nozzles per stage, at least one on the male rotor side and at least one on the female rotor side, and each pair of nozzles is positioned to open simultaneously to a corresponding moving male compression chamber and a moving female compression chamber in the compression chamber.

11. A screw compressor with split flow assisted compression and pulsation traps, comprising: a compression chamber and a pair of meshed multi-spiral-vane rotors housed in the compression chamber, wherein the compression chamber has a suction port and a discharge port, wherein the rotors rotate in the compression chamber to form a series of moving chambers in the compression chamber for sucking and compressing gas and pushing the gas from the suction port toward the discharge port; and

12. The screw compressor according to claim 11 wherein the first stage gas injection nozzle is located at least one rotor pitch from suction (i.e., completely isolated from suction) but before the discharge.

13. The screw compressor of claim 11, further comprising a second stage gas nozzle spaced at least one rotor pitch from the first stage gas nozzle (i.e., completely isolated from the first stage gas nozzle), but before the discharge port, thereby defining a second stage of a feedback flow loop.

14. The screw compressor of claim 11, further comprising a third stage gas nozzle spaced at least one rotor pitch from the second stage gas nozzle (i.e., completely isolated from the second stage gas nozzle), but before the discharge port, thereby defining a third stage of a feedback flow loop.

15. The screw compressor according to claim 11, wherein each stage of gas nozzles is circular in cross-sectional shape with a converging or converging-diverging distribution of cross-sectional area that gradually transitions along the nozzle axis.

16. The screw compressor according to claim 11, wherein each stage of the gas nozzle has a rectangular cross-sectional shape with a converging or converging-diverging distribution of cross-sectional areas that gradually transition along the nozzle axis.

17. The screw compressor according to claim 15, wherein the converging cross-section continuously transitions from a circular cross-sectional shape beginning at the nozzle throat to a rectangular slot shape of an inner wall surface of the compression chamber of equal area, wherein the long sides of the rectangular slot shaped nozzle at the inner wall surface of the compression chamber are parallel to the slot shaped long sides of the moving compression pocket.

18. The screw compressor according to claim 15, wherein the converging-diverging section continuously transitions from a circular cross-sectional shape starting at the nozzle throat to a rectangular slot shape of an inner wall surface of the compression chamber with a gradually diverging area, wherein the long sides of the rectangular slot-shaped nozzle at the inner wall surface of the compression chamber are parallel to the long sides of the slot shape of the moving compression chamber.

19. The screw compressor of claim 11, wherein the stage gas nozzles are positioned a distance away from the rotor axis and point in a tangential direction substantially the same as the direction of rotation of one of the rotors.

20. The screw compressor according to claim 11 wherein the pair of enmeshed multi-lobe rotors comprises a male rotor and a female rotor, with at least two gas injection nozzles per stage, at least one on the male rotor side and at least one on the female rotor side, and each pair of nozzles is positioned to open simultaneously to a corresponding moving male compression chamber and a moving female compression chamber in the compression chamber.