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Publication numberUS20050194207 A1
Publication typeApplication
Application numberUS 10/793,130
Publication dateSep 8, 2005
Filing dateMar 4, 2004
Priority dateMar 4, 2004
Also published asWO2005093246A1
Publication number10793130, 793130, US 2005/0194207 A1, US 2005/194207 A1, US 20050194207 A1, US 20050194207A1, US 2005194207 A1, US 2005194207A1, US-A1-20050194207, US-A1-2005194207, US2005/0194207A1, US2005/194207A1, US20050194207 A1, US20050194207A1, US2005194207 A1, US2005194207A1
InventorsPaul Nemit, James Rizer, Angela Comstock
Original AssigneeYork International Corporation
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Apparatus and method of sound attenuation in a system employing a VSD and a quarter-wave resonator
US 20050194207 A1
Abstract
A resonator in a closed fluid system includes a body having a passageway in communication with a closed fluid system wherein a piston is movable within the passageway. Each position of the piston within the passageway defines a noise attenuation frequency corresponding to a noise frequency generated by the closed fluid system. A device providing a pressurized fluid from the closed fluid system selectively moves the piston within the passageway to vary the noise attenuation frequency.
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Claims(39)
1. A resonator arrangement for a closed fluid system comprising:
a body having a passageway in fluid communication with a closed fluid system;
a piston movable within the passageway wherein a position of the piston within the passageway defines a noise attenuation frequency for the closed fluid system;
a control arrangement to selectively position the piston within the passageway to generate a noise attenuation frequency corresponding to a noise frequency generated by the closed fluid system.
2. The resonator of claim 1 wherein the control arrangement includes a pressurized fluid to selectively position the piston.
3. The resonator of claim 2 wherein the pressurized fluid to selectively position the piston is taken from a pressurized fluid source that is independent of the closed fluid system.
4. The resonator of claim 3 wherein the pressurized fluid to selectively position the piston is separated from the fluid in the closed fluid system.
5. The resonator of claim 3 wherein the pressurized fluid to selectively position the piston is the same as the fluid in the closed fluid system.
6. The resonator of claim 2 wherein the pressurized fluid to selectively position the piston is taken from the closed fluid system.
7. The resonator of claim 3 wherein the pressurized fluid provided to selectively position the piston within the passageway in a direction to generate a decreased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system is taken from a pressurized fluid source that is independent of the closed fluid system; and
wherein the pressurized fluid provided to selectively position the piston within the passageway in a direction to generate an increased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system is taken from the closed fluid system.
8. The resonator of claim 7 wherein the pressurized fluid taken from the pressurized fluid source that is independent of the closed fluid system is separated from the fluid in the closed fluid system.
9. The resonator of claim 3 wherein the pressurized fluid provided to selectively move the piston within the passageway in a direction to generate an increased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system is taken from a pressurized fluid source that is independent of the closed fluid system; and
wherein the pressurized fluid provided to selectively move the piston within the passageway in a direction to generate a decreased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system is taken from the closed fluid system.
10. The resonator of claim 9 wherein the pressurized fluid taken from the pressurized fluid source that is independent of the closed fluid system is separated from the fluid in the closed fluid system.
11. The resonator of claim 1 wherein the closed fluid system is an HVAC&R system.
12. The resonator of claim 11 wherein the HVAC&R system includes a variable speed drive.
13. The resonator of claim 1 wherein the control arrangement includes a valve.
14. The resonator of claim 1 wherein the control arrangement includes a proportional valve.
15. The resonator of claim 1 further comprising a resilient device within the passageway to urge the piston to move in a direction within the passageway.
16. The resonator of claim 15 wherein the resilient device is a spring.
17. The resonator of claim 15 further comprising a resilient device within the passageway to urge the piston to move in a direction within the passageway toward the closed fluid system.
18. The resonator of claim 15 further comprising a resilient device within the passageway to urge the piston to move in a direction within the passageway away from the closed fluid system.
19. A variable noise attenuation device for use with an HVAC&R fluid system comprising:
a body having a passageway in fluid communication with an HVAC&R fluid system;
a piston movable within the passageway wherein a position of the piston within the passageway defining a noise attenuation frequency for the HFAC&R fluid system;
a proportional valve selectively providing a pressurized fluid from the HVAC&R fluid system to the piston to selectively position the piston within the passageway to generate a noise attenuation frequency corresponding to a noise frequency generated by the HFAC&R fluid system.
20. The variable noise attenuation device of claim 19 wherein the proportional valve is in fluid communication with a positive displacement compressor.
21. The variable noise attenuation device of claim 20 wherein the positive displacement compressor is a screw compressor or a reciprocating compressor.
22. The variable noise attenuation device of claim 19 wherein the HVAC&R fluid system is a closed HVAC&R fluid system.
23. The variable noise attenuation device of claim 19 wherein the HVAC&R fluid system includes a variable speed drive.
24. A method for attenuating noise in a closed fluid system, the steps comprising:
providing a body having a passageway in fluid communication with a closed fluid system;
providing a piston movable within the passageway wherein a position of the piston within the passageway defines a noise attenuation frequency for the closed fluid system;
providing a control arrangement for providing a pressurized fluid to the piston; and
positioning selectively the piston within the passageway to generate a noise attenuation frequency corresponding to a noise frequency generated by the closed fluid system.
25. The method of claim 24 wherein the step of providing the control arrangement for providing the pressurized fluid to the piston includes providing the pressurized fluid from a pressurized fluid source that is independent of the closed fluid system.
26. The method of claim 25 wherein the step of providing the control arrangement for providing the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system further includes the step of separating the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system from the fluid in the closed fluid system.
27. The method of claim 25 wherein the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system is the same as the fluid in the closed fluid system.
28. The method of claim 24 wherein the step of providing the control arrangement for providing the pressurized fluid to the piston includes providing the pressurized fluid taken from the closed fluid system.
29. The method of claim 25 wherein the step of providing the control arrangement for providing the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system to the piston includes providing the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system to selectively position the piston within the passageway in a direction to generate a decreased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system; and
providing the control arrangement for providing the pressurized fluid from the closed fluid system to selectively position the piston within the passageway in a direction to generate an increased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system.
30. The method of claim 29 wherein the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system to selectively position the piston within the passageway in a direction to generate the decreased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system is separated from the fluid in the closed fluid system.
31. The method of claim 25 wherein the step of providing the control arrangement for providing the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system to the piston includes providing the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system to selectively position the piston within the passageway in a direction to generate an increased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system; and
providing the pressurized fluid from the closed fluid system to selectively position the piston within the passageway in a direction to generate a decreased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system.
32. The method of claim 31 wherein the pressurized fluid from the pressurized fluid source that is independent of the closed fluid system to selectively position the piston within the passageway in a direction to generate an increased noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system is separated from the fluid in the closed fluid system.
33. The method of claim 24 wherein the closed fluid system is an HVAC&R system.
34. The method of claim 33 wherein the HVAC&R system includes a variable speed drive.
35. The method of claim 24 wherein the step of providing the control arrangement includes a step of providing a valve.
36. The method of claim 24 wherein the step of providing the control arrangement includes providing a proportional valve.
37. The method of claim 24 further including an additional step, before the step of providing a control arrangement for providing a pressurized fluid from the closed fluid system to the piston, of providing an algorithm, the step of providing an algorithm further including the steps of
reading a variable speed drive frequency output; and
comparing the variable speed drive frequency output to the position of the piston in the passageway; and
wherein the step of positioning selectively the piston within the passageway to generate the noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system includes positioning selectively the piston within the passageway to generate the noise attenuation frequency corresponding to the noise frequency generated by the closed fluid system in response to the algorithm.
38. The method of claim 24 further including an additional step, before the step of providing a control arrangement for providing pressurized fluid, of providing a resilient device within the passageway to urge the piston into movement within the passageway toward the closed fluid system.
39. The method of claim 38 wherein the resilient device is a spring.
Description
FIELD OF THE INVENTION

The present invention relates generally to a method of operation and apparatus for noise attenuation of closed fluid systems, and more particularly, to a method of operation and apparatus for adjustable noise attenuation of closed HVAC&R systems having positive displacement compressors with variable speed drives.

BACKGROUND OF THE INVENTION

Heating, ventilation, air conditioning or refrigeration (HVAC&R) systems typically maintain temperature control in a structure by circulating a fluid within coiled tubes such that passing another fluid over the tubes effects a transfer of thermal energy between the two fluids. A primary component in such a system is a positive displacement compressor, which receives a cool, low pressure gas and by virtue of a compression device, exhausts a hot, high pressure gas. One type of positive displacement compressor is a screw compressor, which generally includes two cylindrical rotors mounted on separate shafts inside a hollow, double-barreled casing. The side-walls of the compressor casing typically form two parallel, overlapping cylinders which house the rotors side-by-side. Screw compressor rotors typically have helically extending lobes and grooves on their outer surfaces forming a large thread on the circumference of the rotor. During operation, the threads of the rotors mesh together, with the lobes on one rotor meshing with the corresponding grooves on the other rotor to form a series of gaps between the rotors. These gaps form a continuous compression chamber that continuously reduces in volume as the rotors turn to compress the gas and which communicates with the compressor inlet opening, or “port,” at one end of the casing and with a discharge port at the opposite end of the casing.

These rotors rotate at high rates of speed, and multiple sets of rotors (compressors) may be configured to work together to further increase the amount of gas that can be circulated in the system, thereby increasing the operating capacity of a system. While the rotors provide a continuous pumping action, each set of rotors (compressor) produces pressure pulses as the pressurized fluid is discharged at the discharge port. These discharge pressure pulsations act as significant sources of audible sound within the system. In addition, mechanical noise is generated by the meshing and unmeshing of the rotors which is the result of backlash and movement between the rotors.

To eliminate or minimize the undesirable sound, noise attenuation devices or systems can be used. Noise attenuation devices include reflective and absorptive mufflers. Absorptive mufflers are most effective when the frequencies of the sound are greater than 500 Hz. Since the fundament frequencies for a positive displacement compressor, such as a screw compressor, are typically less than 500 Hz, absorptive mufflers are less effective than other types of mufflers.

However, the pressure pulse frequency of screw compressors are predictable, that is, the pressure pulse frequency is a function of the number of lobes of the male rotor multiplied by its rotational speed. For example, a screw compressor having a 5 lobed male rotor, which rotates at 3,600 RPM generates a frequency of 5×3, 600/60 or 300 Hz. Therefore, for a compressor, such as a screw compressor, operating at a fixed speed, a reflective muffler, such as a side branch resonator, which is also referred to as a quarter wave resonator, may be configured to attenuate noise at a specific frequency. The side branch resonator typically comprises a component that forms a tee “T” adjacent a compressor discharge, with the tee being closed at the end opposite the compressor discharge. The length of the side branch resonator is configured or “tuned” to be one-fourth of one wavelength of the sound produced by the compressor to achieve the desired noise attenuation.

Unfortunately, even with the compressor operating at a fixed speed, there are variations in ambient conditions, such as pressure and temperature of the refrigerant gas used in the system. Variations in the ambient conditions can likewise vary the frequency of the noise produced by the system, as such ambient condition variations can likewise change the acoustic velocity of the refrigerant gas. The term acoustic velocity is defined as the velocity that sound travels through the refrigerant gas for a given set of ambient conditions. Thus, it is generally not possible to attenuate the noise generated by a fixed speed compressor with a single side branch or quarter-wave resonator. Further, for reasons of increased compressor efficiency, it is often desirable to employ variable speed drives to power the compressor motors, resulting in compressors running at variable speeds.

What is needed is a cost-effective, efficient and easily implemented method or apparatus for compressor noise attenuation that may be used with variable speed compressors.

SUMMARY OF THE INVENTION

The present invention relates to a resonator arrangement for a closed fluid system including a body having a passageway in fluid communication with a closed fluid system. A piston is movable within the passageway wherein a position of the piston within the passageway defines a noise attenuation frequency for the closed fluid system. A control arrangement selectively positions the piston within the passageway to generate a noise attenuation frequency corresponding to a noise frequency generated by the closed fluid system.

The present invention further relates to a variable noise attenuation device for use with an HVAC&R fluid system including a body having a passageway in fluid communication with an HVAC&R fluid system. A piston is movable within the passageway wherein a position of the piston within the passageway defines a noise attenuation frequency for the HVAC&R fluid system. A proportional valve selectively provides a pressurized fluid from the HVAC&R fluid system to the piston to selectively position the piston within the passageway to generate a noise attenuation frequency corresponding to a noise frequency generated by the HVAC&R fluid system.

The present invention yet further relates to a method for attenuating noise in a closed fluid system having the steps of providing a body having a passageway in fluid communication with a closed fluid system; providing a piston movable within the passageway wherein a position of the piston within the passageway defines a noise attenuation frequency for the closed fluid system; providing a control arrangement for providing a pressurized fluid to the piston; and positioning selectively the piston within the passageway to generate a noise attenuation frequency corresponding to a noise frequency generated by the closed fluid system.

An advantage of the present invention is that it attenuates noise of variable frequency.

A further advantage of the present invention is that it is inexpensive to manufacture.

Other features and advantages of the present invention will be apparent from the following more detailed description of the preferred embodiment, taken in conjunction with the accompanying drawings which illustrate, by way of example, the principles of the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic of a variable speed HVAC&R system having a variable frequency resonator of the present invention.

FIG. 2 is an enlarged partial schematic of the HVAC&R system including an enlarged elevation view of the variable frequency resonator of the present invention.

FIG. 3 is an enlarged partial schematic of the HVAC&R system including an enlarged elevation view of an embodiment of the variable frequency resonator of the present invention.

Wherever possible, the same reference numbers will be used throughout the drawings to refer to the same or like parts.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1 illustrates generally a HVAC&R system 10 incorporating the present invention. An AC power source 12 supplies AC power to a variable speed drive (VSD) 14, which in turn, supplies AC power to a motor 18 for driving a compressor 20. A control panel 16 controls both the output frequency and the voltage from VSD 14 that is supplied to motor 18, which selectively varies both the torque and speed of motor 18, and likewise varies the speed of compressor 20. Control panel 16 includes an analog to digital (A/D) converter, a microprocessor, a non-volatile memory, and an interface board to control operation of the refrigeration system 10. The control panel 16 can also be used to control the operation of the VSD 14, the monitor 18 and the compressor 20. Compressor 20 receives refrigerant gas at its inlet and discharges compressed refrigerant gas from its outlet through discharge line 30. Optionally, other compressors may operate in parallel with compressor 20 to increase the capacity of the system 10. These compressors are typically positive displacement compressors, such as screw, reciprocating or scroll, having a wide range of cooling capacity, but may also include other compressor constructions.

After refrigerant gas is discharged by compressor 20 through discharge line 30 toward condenser 22, the refrigerant gas is directed past a variable resonator 26, such as a quarter wave resonator. Variable resonator 26 is continuously tuned, as will be described in further detail below, to attenuate both the pressure pulses and the mechanical noise generated by the meshing and unmeshing of the rotors that is the result of backlash and movement between the rotors. The variable resonator 26 can attenuate both the pressure pulses and the mechanical rotor noise, since each occurs at substantially the same frequency. Once the refrigerant gas passes variable resonator 26, it is serially directed to condenser 22 and then to evaporator 24 where the refrigerant gas is placed in a non-mixing heat exchange relationship with fluids outside the system 10 to maintain temperature control in a structure before returning the refrigerant gas to the compressor 20 to repeat the cycle. The conventional HVAC&R system 10 includes many other features that are not shown in FIG. 1. These features have been purposely omitted to simplify the drawing for ease of illustration.

Referring to FIG. 2, variable resonator 26 includes a body 38, such as a tube or cylinder, which defines a passageway 39. A stop 48 having an aperture 49 is secured to one end of body 38, and a cap 40 is secured to the opposite end of body 38. The end of body 38 having the stop 48 is connected to discharge line 30 so that passageway 39 of body 38 is in fluid communication with the refrigerant gas flowing through discharge line 30. A piston 42 is inserted inside of passageway 39 that is slidable or movable within passageway 39, the piston 42 further includes a seal 44, such as an O-ring, to provide a substantially fluid tight seal between piston 42 and an inside surface 41 of body 38 as piston 42 slides along passageway 39. A resilient device 46, such as a spring, is interposed between piston 42 and cap 40 to urge piston 42 into movement along passageway 39 in a direction away from cap 40, the travel of the piston 42 in this direction being limited by stop 48. Thus, the travel of the piston 42 along passageway 39 is limited by the stop 48, in one direction, and by the cap 40, minus the thickness of the compressed resilient device 46, in the other direction.

It is desirable to control the position of the piston 42 within the passageway 39, in which a variable resonator length 50 is defined by the distance between the lower surface of piston 42 and the lower surface of stop 48. The variable resonator length 50 of resonator 26, which is a side branch resonator, operates in accordance with the following equation stated symbolically as
F=C/(4×1)  [1]

Wherein F is the resonant frequency of the resonator 26; C is the sonic speed or acoustic velocity of the fluid traveling through the resonator 26; and 1 is the resonator length 50. Thus, the variable resonator length 50 preferably corresponds to one quarter of a wavelength, such as the wavelengths generated during the operation of the compressor 20, which can be attenuated by the resonator 26. While the single position of a fixed length resonator may effectively attenuate noise of a specific wavelength corresponding to a compressor operating at a fixed speed and a specific set of ambient conditions, such unchanging operating conditions are unrealistic. In other words, changing ambient conditions, such as temperature, also changes the density of the refrigerant, which changes the sonic speed C of the refrigerant, which changes the resonator length required to attenuate the noise generated by the compressor as discussed in equation [1]. A fixed length resonator cannot accommodate these changing conditions. Additionally, the compressor of the present invention operates at variable speeds, which, in turn, results in variable frequencies. Thus, selective control of the position of the piston 42 within the passageway 39 of the resonator 26 is required for consistent, noise attenuation.

To achieve positional control of the piston 42 within the passageway 39, a valve 28, such as a proportional valve, is controlled by the control panel 16 to selectively provide pressurized refrigerant gas from a pressure line 36 through cap 40. To increase the amount of pressure in the pressure line 36, the control panel 16 actuates the valve 28 to bleed pressurized refrigerant gas from a high pressure line 34, which is in fluid communication with the compressor discharge 30, to provide to the pressure line 36. Conversely, to decrease the amount of pressure in the pressure line 36, the control panel 16 actuates the valve 28 to bleed pressurized refrigerant gas from the pressure line 36 to provide to the low pressure line 32, which may be connected to a port in the compressor 20, or to the suction line at the inlet of the compressor 20. Alternately, the low pressure line 32 may be connected to a separate container (not shown) that is maintained at an even lower pressure than the suction line at the compressor inlet to provide an even wider range of pressure differential, if desired. In other words, the amount of pressure in the pressure line 36 can range anywhere from the discharge pressure of the high pressure line 34 to the lower pressure contained in the low pressure line 32 to effect movement of the piston 42 along the passageway 39 of the resonator body 38.

To effect movement of the piston 42 within the passageway 39, the net force that is applied to the surface of the piston 42, which is facing the cap 40 must be sufficiently different from the net force that is applied to the opposite surface of the piston 42 which faces the aperture 49. There must be a sufficient difference in forces acting on the opposed surfaces of the piston 42 to overcome any combination of frictional forces between the seal 44 and the inside surface of body 38, the weight of the piston 42 (if the piston moves vertically) and inertia of the piston 42. The net force that is applied to the surface of the piston 42 facing the cap 40 is applied in two parts. The first part of the net force is the pressure of the refrigerant gas in the pressure line 36 (PLINE) multiplied by the surface area of the piston 42 (ASURFACE) facing the cap 40. The second part of the net force is the force applied by the resilient device 46, which is the product of the spring constant (k) of the spring multiplied by the distance (d) the spring is compressed. The sum of the forces applied to the surface of piston 42 facing cap 40 (FCAP) is stated symbolically as
F CAP=(P LINE ×A SURFACE)+(k×d)  [2]

The net force that is applied to the surface of the piston 42 facing the stop 48 is applied in a single part. This single net force part is the pressure of the refrigerant gas in the discharge line 36 (PDISCHARGE) multiplied by the surface area of the piston 42 (ASURFACE) facing the stop 48. The force applied to the surface of the piston 42 facing the stop 48 (FSTOP) is stated symbolically as
F STOP =P DISCHARGE ×A SURFACE  [3]

Depending upon the configuration of the HVAC&R system, the PDISCHARGE may or may not vary. For example, if the system uses a water-cooled condenser, PDISCHARGE is considered substantially constant over the range of operating speeds of the compressor. In contrast, if an air-cooled condenser is used, the PDISCHARGE fluctuates, the fluctuation being either variable or substantially constant, depending upon the operating conditions. However, over an extreme range of operating conditions using the air-cooled condenser, it has been determined that the magnitude of PDISCHARGE will vary no more than about one percent of the variable resonator length 50. Thus, the effect of PDISCHARGE fluctuation can typically be disregarded, while it is the sonic speed of the refrigerant that must be monitored.

Therefore, when there is a sufficient difference between the FCAP and the FSTOP forces, the piston 42 is urged to move along the passageway 39 in a direction away from the larger force. In other words, the piston 42 will continue to move in this direction until the forces are substantially equal, or the piston 42 has reached the end of its possible travel, that is, abutting the stop 48 or abutting the cap 40, minus the thickness of the compressed resilient device 46. When the piston 42 abuts stop 48, the variable resonator length 50 is at its shortest length and corresponds to the shortest one-quarter wavelength position, and therefore, the shortest wavelength, or highest frequency, that the resonator 26 can attenuate. Conversely, when the piston 42 abuts the cap 40, minus the thickness of the compressed resilient device 46, the variable resonator length 50 is at its maximum length, and corresponds to the longest one-quarter wavelength position, and therefore, the longest wavelength, or lowest frequency, that the resonator 26 can attenuate. It is appreciated by those skilled in the art that there exists a reciprocal relationship between frequency and wavelength.

The position of the piston 42 is controlled by the control panel 16 using a control algorithm. Based upon application of equations [1]-[3], the control algorithm determines the desired resonator length 50, that is, the desired position of the piston 42 within the passageway 39, by reading the VSD frequency output that is supplied to the compressor motor 18, and comparing the desired position of the piston 42 to the position of the piston 42 in the passageway 39. The control algorithm may employ a look-up table that accounts for changes in ambient conditions, such as temperature and pressure, which likewise can change the acoustic velocity of the gas, and the desired position of the piston 42 within the passageway 39. Further, the control algorithm may further employ the look-up table or calculate the relationship between the spring length (d in equation [2]) and the piston 42 position without needing to measure the piston 42 position. Alternatively, the position of the piston 42 in the passageway 39 may be determined by a sensor, a rheostat, or any other suitable mechanical or electrical device that provides such positional information to the control panel 16. In response to comparing the desired position of the piston 42 to its actual position within the passageway 39, the control panel 16 sends a control signal(s) to the valve 28. If the control algorithm determines that the variable resonator length 50 should be decreased, the control signal from the control panel 16 causes the valve 28 to actuate to bleed pressurized refrigerant from the high pressure line 34 to the pressure line 36, thereby increasing the pressure in the pressure line 36. The combined forces, generated from the pressure in the pressure line 36 and the compressed resilient device 46, are applied to the piston 42, which collectively define force FCAP as previously discussed in equation [2]. The opposing force that is applied to the piston 42 defines force FSTOP as previously discussed in equation [3]. When the FCAP force sufficiently exceeds the FSTOP force, the piston 42 slides toward the stop 48, thereby decreasing the variable resonator length 50.

Conversely, if the algorithm determines that the resonator length 50 should be increased, the control signal from the control panel 16 causes the valve 28 to actuate to bleed pressurized refrigerant from the pressure line 36 to the low pressure line 32, thereby decreasing the pressure in the pressure line 36. The combined forces generated from the pressure in the adjustable pressure line 36 and the compressed resilient device 46 are applied to the piston 42, which collectively define force FCAP as previously discussed in equation [2]. The opposing force that is applied to the piston 42 defines force FSTOP as previously discussed in equation [3]. When the FSTOP force sufficiently exceeds the FCAP force, the piston 42 slides toward the cap 40, thereby increasing the variable resonator length 50.

Although a preferred embodiment of the resonator 26 discloses the valve 28 controlling the position of the piston 42 by selectively providing pressurized refrigerant gas between lines 32, 34 and 36 using the refrigerant gas from the HVAC&R system 10, it is appreciated that pressurized refrigerant from a source that is independent from the HVAC&R system 10 may also be used. Additionally, other pressurized fluids each from independent sources of pressurized fluid such as air or other compressible gas, or even incompressible fluids may also be used with the valve 28 to selectively move the piston 42 along the passageway 39. However, the use of such gases or fluids, if incompatible with the efficient operations HVAC&R system 10, may need to remain separate from the refrigerant in the HVAC&R system 10. Alternately, piston 42 may also be magnetically or electrically displaced within the passageway 39.

For example, referring to FIG. 3, which is otherwise the same as FIG. 2 except as shown, pressurized fluids are selectively provided to or bled from the pressure line 36 of a resonator 126 by the valve 28 from an independent pressurized fluid source 154 and an independent fluid receptacle 152. The fluid receptacle 152 receives pressurized fluid that is bled from the pressure line 36 when the control panel 16 determines that the variable resonator length 50 needs to be increased to likewise urge the piston 42 to move along the passageway 39 of the resonator body 38 toward the cap 40. Conversely, the pressurized fluid source 154 provides pressurized fluid to the pressure line 36 when the control panel 16 determines that the variable resonator length 50 needs to be decreased to likewise urge the piston 42 to move along the passageway 39 of the resonator body 38 toward the stop 48. The pressurized fluid source 154 is pressurized to a higher level than the receptacle 152, and may be pressurized to a much higher pressure than the discharge pressure from the discharge line 30, if desired. By providing highly pressurized fluid in pressurized fluid source 154, it may be possible to eliminate the need for the resilient device spring 46, and decrease the response time for movement of the piston 42 along the passageway 39 of the resonator body 38.

Optionally, a second valve 129 that is also controlled by the control panel 16 is provided to selectively provide pressurized fluid from a high pressure line 158 to a variable pressure line 156. To decrease the amount of pressure in the variable pressure line 156, pressurized fluid may be bled to the receptacle 152 through a low pressure line 160. To ensure the pressurized fluid that is provided to the resonator body 38 from the second valve 129 remains separate from the refrigerant fluid of the HVAC&R system 10, i.e., the refrigerant gas flowing in the discharge line 30, the pressurized fluid from the second valve 129 is received in an expandable/contractable toroidal container 162. The container 162 is disposed between the stop 48 and the piston 42 and is composed of a resilient material that permits the container 162 to expand or contract in response to sufficient differences between forces FCAP and FSTOP which result in movement of the piston 42 with respect to the container 162. Preferably, the container 162 remains adjacent to the inside the surface 41 of the resonator body 38 and collapses upon itself as the container 162 is maintained in contact with the piston 42 as it moves along the passageway 39, possibly by use of baffles. Optionally, a resilient device (not shown) may be provided inside the container 162. Additionally, the container 162 preferably provides a uniform cross sectional profile within the passageway 39 to further provide predictable noise attenuation behavior as the resonator length 50 is varied as previously described.

It is appreciated that any number of combinations of independently provided pressurized fluids from pressurized sources or receptacles may be used with the pressurized refrigerant gas from the HVAC&R system, and that these pressurized fluids may or may not need to be separated from each other. It is also appreciated that any number of valves may be used to control the flow of the pressurized fluids as discussed.

While the invention has been described with reference to a preferred embodiment, it will be understood by those skilled in the art that various changes may be made and equivalents may be substituted for elements thereof without departing from the scope of the invention. In addition, many modifications may be made to adapt a particular situation or material to the teachings of the invention without departing from the essential scope thereof. Therefore, it is intended that the invention not be limited to the particular embodiment disclosed as the best mode contemplated for carrying out this invention, but that the invention will include all embodiments falling within the scope of the appended claims.

Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US7757808 *Feb 4, 2009Jul 20, 2010Gm Global Technology Operations, Inc.Noise reduction system
US7946382 *May 23, 2007May 24, 2011Southwest Research InstituteGas compressor with side branch absorber for pulsation control
US8123498Jan 24, 2008Feb 28, 2012Southern Gas Association Gas Machinery Research CouncilTunable choke tube for pulsation control device used with gas compressor
US8393437 *Feb 15, 2011Mar 12, 2013Westinghouse Electric Company LlcNoise and vibration mitigation system for nuclear reactors employing an acoustic side branch resonator
US8444397 *Sep 7, 2011May 21, 2013Johnson Controls Technology CompanyManual selective attenuator
US20120206011 *Feb 15, 2011Aug 16, 2012Westinghouse Electric CompanyNoise and vibration mitigation system for nuclear reactors employing an acoustic side branch resonator
US20130028758 *Sep 7, 2011Jan 31, 2013Johnson Controls Technology CompanyManual selective attenuator
US20130306398 *May 15, 2013Nov 21, 2013Leica Microsystems Cms GmbhApparatus for Damping Sound in the Optical Beam Path of a Microscope, and Microscope Having a Corresponding Apparatus
WO2011152915A2Mar 16, 2011Dec 8, 2011Carrier CorporationPulsation cancellation
Classifications
U.S. Classification181/250, 181/276
International ClassificationF02M35/12, B60H1/00
Cooperative ClassificationF02M35/1222, F02M35/1255, F25B1/04, F25B2500/12, F25B49/00
European ClassificationF25B49/00, F02M35/12
Legal Events
DateCodeEventDescription
Mar 4, 2004ASAssignment
Owner name: YORK INTERNATIONAL CORPORATION, PENNSYLVANIA
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNORS:NEMIT, PAUL, JR.;RIZER, JAMES BRIAN;COMSTOCK, ANGELA MARIE;REEL/FRAME:015052/0109
Effective date: 20040303