Patent Application
20070028647
The present invention is directed to a refrigerant flow control and pressure recovery device for use with vapor compression refrigeration systems in HVAC applications, and more particularly, is directed to a refrigerant pressure recovery device for use with shell and tube type condensers, where cooling fluids such as water flows through tubes and the refrigerant flows through the shell and is cooled and condensed on the outside surface of the tubes.
Condensers are an important component used in vapor compression refrigeration systems in HVAC applications. In one type of condenser, refrigerant vapor enters the shell of the condenser and flows across the outside surface of a plurality of cooling tubes. Each of the tubes contains a cooling fluid (e.g., water) at a lower temperature circulating inside these tubes. As the refrigerant flows outside of the tubes, heat transfer occurs from the refrigerant to the lower temperature fluid circulating inside the tubes, such that refrigerant temperature lowers below the saturation temperature and condenses on the outside of the tubes. The condensed refrigerant exits the condenser in the liquid state and the warmer fluid circulating inside the tubes is typically directed to a cooling tower. The condensed liquid refrigerant from the condenser flows through an expansion device to an evaporator. The two-phase refrigerant in the evaporator enters into a heat exchange relationship with a secondary fluid to lower the temperature of the secondary fluid that is circulated to regulate the temperature of an area inside a structure. The refrigerant liquid in the evaporator undergoes a phase change to a refrigerant vapor as a result of the heat exchange relationship with the secondary liquid and is returned to the compressor where the pressure of the refrigerant vapor is elevated and discharged into the condenser to complete the cycle.
In a typical embodiment of the above system, the refrigerant vapor from the discharge of a compressor enters the shell of the condenser at relatively high velocity. An impingement baffle is typically disposed on the inlet of a shell side condenser to prevent direct impingement of the high velocity refrigerant vapor on the condenser tubes. This direct impingement can cause damage to the condenser tubes, such as by vibration, pitting and erosion. Conventional impingement baffles define an elongate, narrow chamber that directs the incoming refrigerant vapor toward opposed ends of the condenser. While preventing damage to the condenser tubes, the impingement baffle causes a drop in pressure of the incoming refrigerant vapor as compared to the pressure of the refrigerant vapor at the compressor outlet. The compressor needs to compress the refrigerant vapor to a higher pressure to make up this pressure drop with more power consumption, thereby lowering the overall refrigeration system efficiency. In addition, a portion of the incoming refrigerant vapor traveling along the center of the baffle retains its high velocity, resulting in potential tube vibration problems and a phenomenon referred to as “liquid hump.” Liquid hump refers to a rise in the level of the condensed refrigerant liquid in the central portion of the condenser shell as compared to the level at the ends of the condenser shell thereby reducing the effective heat transfer surface area, which can reduce condenser efficiency. Further, the high velocity refrigerant causes undesirable splashing of the liquid refrigerant in the condenser shell.
What is needed is a diffuser at the condenser inlet that smoothly decelerates and transitions incoming refrigerant flow, achieving minimum stagnation pressure losses and maximizes the static pressure recovery from the inlet of the condenser to the exit of the diffuser (which is the pressure inside the shell). For a given condensing saturation pressure and temperature inside the shell, a lower pressure is needed at the inlet to the condenser when using the diffuser of the present invention, thereby reducing the power consumption of the compressor, and thereby improving the system efficiency. In the conventional condenser, there is a drop in pressure across the impingement plate as opposed to the case with the diffuser of the present invention where a static pressure recovery occurs. Alternatively, if maintaining the same compressor discharge pressure, the condensing saturation pressure and temperature inside the shell is higher when using the diffuser of the present invention due to static pressure recovery. Without altering the temperature of the fluid circulating through the condenser tubes to cool the refrigerant, the temperature difference between the two is increased, so that less heat transfer surface is needed to reject the same amount of heat. Using the diffuser of the present invention provides opportunity to use a smaller condenser to achieve the same system efficiency.
The present invention relates to a diffuser situated at the inlet of a shell side condenser of a vapor compression refrigeration system. The diffuser includes an inlet to receive a compressed refrigerant vapor from a compressor of the vapor compression refrigeration system. A chamber is in fluid communication with the inlet to receive compressed refrigerant vapor. The chamber has an upper side, a lower side and lateral sides bridging the upper and lower sides. The chamber also has a plurality of openings to discharge refrigerant vapor inside the condenser shell. A protrusion is disposed inside the chamber. The protrusion and the chamber are configured and disposed to diffuse and direct a flow of refrigerant from the discharge of the compressor to inside the condenser. The refrigerant leaving the chamber of the diffuser has a higher pressure than the refrigerant entering the diffuser at the inlet of the condenser. The inlet of the diffuser typically is in very close proximity to the compressor discharge.
The present invention further relates to a chiller system including a compressor, a condenser arrangement and an evaporator arrangement connected in a closed refrigerant loop. An inlet is in fluid communication between the compressor and the condenser arrangement to receive a compressed refrigerant vapor from the compressor. A chamber is in fluid communication with the inlet to receive compressed refrigerant. The chamber has an upper side, a lower side and lateral sides bridging the upper and lower sides. The chamber also has a plurality of openings to discharge refrigerant inside the condenser arrangement. A protrusion is disposed inside the chamber. The protrusion and the chamber are configured and disposed to diffuse and direct a flow of refrigerant from the discharge of the compressor to inside the condenser. The refrigerant leaving the chamber has a higher pressure than the refrigerant entering the chamber at the inlet of the condenser. The inlet of the condenser typically is in very close proximity to the compressor discharge.
The present invention still further relates to a condenser including an inlet to receive a compressed refrigerant from a compressor of a vapor compression refrigeration system. A chamber is in fluid communication with the inlet to receive compressed refrigerant. The chamber has an upper side, a lower side and lateral sides bridging the upper and lower sides. The chamber also has a plurality of openings to discharge refrigerant inside the condenser. A protrusion is disposed inside the chamber. The protrusion and the chamber are configured and disposed to diffuse and direct a flow of refrigerant from the discharge of the compressor to inside the condenser. The refrigerant leaving the chamber has a higher pressure than the refrigerant entering the chamber at the inlet of the condenser. The inlet of the condenser typically is in very close proximity to the compressor discharge.
An advantage of the present invention is that it facilitates static pressure recovery of the refrigerant entering the condenser, thereby increasing the pressure of the refrigerant vapor leaving the diffuser compared to the pressure of refrigerant entering the diffuser
An advantage of the present invention is that it increases vapor compression refrigeration system efficiency.
A further advantage of the present invention is that it reduces tube vibration associated with operation of the condenser.
A yet additional advantage of the present invention is that it reduces the level of liquid hump inside the condenser.
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.
Wherever possible, the same reference numbers will be used throughout the drawings to refer to the same or like parts.
One embodiment of a refrigeration system 100 using a shell side condenser inlet diffuser 114 of the present invention is shown in
The evaporator 116 can include connections for a supply line and a return line of a cooling load. A secondary liquid, e.g., water, ethylene or propylene glycol, calcium chloride brine or sodium chloride brine, travels into the evaporator 116 via return line and exits the evaporator 116 via supply line. The liquid refrigerant in the evaporator 116 enters into a heat exchange relationship with the secondary liquid to lower the temperature of the secondary liquid. The refrigerant liquid in the evaporator 116 undergoes a phase change to refrigerant vapor as a result of the heat exchange relationship with the secondary liquid. The vapor refrigerant in the evaporator 116 exits the evaporator 116 and returns to the compressor 110 by a suction line to complete the cycle. It is to be understood that the diffuser 114 is applied to shell side condensers of shell and tube type 112 where the refrigerant condenses on the outside of the tubes (shell side) whereas the evaporator 116 used in the system 100 can be of any suitable configuration, provided that the appropriate phase change of the refrigerant in the condenser 112 and evaporator 116 is obtained.
The refrigeration or liquid chiller system 100 can include many other features that are not shown in
Referring to
In an alternate embodiment, referring to
Similarly, one end of edge 148 is defined by a juncture 144 which is the juncture between edge 148 and edge 154, and preferably, also coincides with the tangency curve 134. The other end of edge 148 is defined by a juncture 170 between edge 148 and edge 150. Preferably, edges 146, 148 each define outwardly directed curves, or convex profiles with regard to lobe 138, although edges 146, 148 can define non-convex profiles, including linear profiles. By edges 146, 148 having convex profiles or suitable non-convex profiles, it can be shown that any line drawn parallel to a reference line 182 connecting junctures 142, 144 along upper surface 138 between junctures 168, 170 is longer than reference line 182. When the diffuser 114 is installed, the reference line 182 is substantially transverse to the length of the condenser 112. The diffuser 114 is preferably bifurcating the flow of refrigerant vapor entering the diffuser 114 along the inlet pipe 127. Stated another way, the distance between corresponding points along edges 146, 148 parallel to reference line 182 increases as the distance of the parallel lines from the reference line 182, i.e., the parallel lines moving along lobe 138 toward edge 150, increases.
Adjacent to edges 146 and 148 and defined by respective junctures 168, 170 is edge 150. Edge 150 is preferably outwardly directed or convex with respect to upper surface 138. Preferably, the curvature of edge 150 is substantially radial, with the center of curvature being coincident with the center of a projection 176. However, the curvature of edge 150 can also be elliptical.
Similar to lobe 138, lobe 140 is defined by edges 152, 154 and 156, while lobe 138 is defined by edges 146, 148 and 150. One end of edge 152 is defined by a juncture 142 which is the juncture between edges 146 and edge 152, and preferably, also coincides with the tangency curve 134. The other end of edge 152 is defined by a juncture 172 between edge 152 and edge 156.
Similarly, one end of edge 154 is defined by a juncture 144 which is the juncture between edge 148 and edge 154, and preferably, also coincides with the tangency curve 134. The other end of edge 154 is defined by a juncture 174 between edge 154 and edge 156. Preferably, edges 152, 154 each define outwardly directed curves, or convex profiles with regard to lobe 140, although edges 152, 154 can define non-convex profiles, including linear profiles. By edges 152, 154 having convex profiles or suitable non-convex profiles, it can be shown that any line drawn parallel to a reference line 182 connecting junctures 142, 144 along lobe 140 and between junctures 172, 174 is longer than reference line 182. Stated another way, the distance between corresponding points along edges 152, 154 parallel to reference line 182 increases as the distance of the parallel lines from the reference line 182, i.e., the parallel lines moving along lobe 140 toward edge 156, increases.
Adjacent to edges 152 and 154 and defined by respective junctures 172, 174 is edge 156. Edge 156 is preferably outwardly directed or convex with respect to lobe 140. Preferably, the curvature of edge 156 is substantially radial, with the center of curvature being coincident with the center of projection 176. However, the curvature of edge 156 can also be elliptical.
Although in a preferred embodiment lobes 138, 140 are symmetrical to each other about the reference line 182 that is preferably coincident with the apex of the protrusion 176, lobes 138, 140 may have a different line of symmetry, lack a line of symmetry, or be asymmetric to each other.
A lower surface 158 is substantially similar in size and shape as upper surface 136, with lower surface 158 and upper surface 136 being separated by a distance 184 that is configured to yield the most favorable results, primarily based on the refrigerant flow rate. Protrusion 176 preferably extends upwardly from the lower surface 158 to help smoothly transition substantially vertically directed refrigerant vapor flow to substantially horizontally directed refrigerant vapor flow upon leaving the diffuser 114. In a preferred embodiment, protrusion 176 is a right circular cone, with the apex of the cone disposed coincident with the center of the neck 128 of the inlet pipe 127. However, it is to be understood that other protrusion geometries can also be used. Further, while protrusion 176 is affixed to the lower surface 158, the protrusion 176 can also be positioned using any suitable mounting arrangement in the refrigerant vapor flow stream between the upper surface 136 and the lower surface 158, or if the protrusion is large enough in at least one direction, to be positioned between the lower surface 158 and the inlet tube 127.
Extending between and bridging the upper and lower surfaces 136, 158 are lateral surfaces 160, 162, 164, 166. Preferably, lateral surface 160 bridges the upper and lower surfaces 136, 158 between juncture 142 and juncture 168 and lateral surface 164 bridges the upper and lower surfaces 136, 158 between juncture 142 and juncture 172. It is similarly preferred that lateral surface 162 bridges the upper and lower surfaces 136, 158 between juncture 144 and juncture 170 and lateral surface 166 bridges the upper and lower surfaces 136, 158 between juncture 144 and juncture 174. In other words, refrigerant vapor that is directed inside the inlet pipe 127, through the flared portion 132, then between the upper and lower surfaces 136, 158 is substantially constrained to flow through an opening 186 between corresponding edges 150 of the upper and lower surfaces 136, 158 in one direction, and an opening 188 between corresponding edges 156 of the upper and lower surfaces 136, 158 in the other direction.
Referring to
Stated another way, refrigerant vapor flowing inside inlet tube 127 past the flared portion 134 and between upper and lower surfaces 136, 158 impinges upon the projection or protrusion 176, which conditions the flow of the refrigerant vapor from a substantially vertical direction to a substantially horizontal direction. The refrigerant vapor is then additionally constrained to flow inside the upper and lower surfaces 136, 158 and lateral surfaces 160, 162, 164, 166 toward the opposed ends 150, 156, the cross sectional area defined by these surfaces increasing as the refrigerant vapor flows toward the opposed ends 150, 156.
By virtue of the ever-increasing cross sectional area along the diffuser surfaces, the projection 176 and the flared portion 134, the vapor refrigerant flow is advantageously conditioned and controlled. That is, the flow of refrigerant vapor is turned substantially 90 degrees while keeping flow losses at a minimum.
Before further analyzing the condenser inlet of the present invention, a brief discussion is provided by a modified form of Bernoulli's equation as provided in Equation 1 below, which can serve as an intuitive guide for analyzing such flows:
P1/(ρg)+(1/(2g))(U1)2=P2/(ρg)+(1/(2g))(U2)2+Loss [1]
wherein P1, P2 is the pressure at a location 1 and a location 2, respectively, ρ is the density of the flowing fluid, U1, U2 is the velocity of the flowing fluid at a location 1 and a location 2, respectively. The (1/(2 g))(U)2 term represents the kinetic energy component also referred to as velocity head of the fluid at locations 1 and 2, respectively. The P/(ρ g) component of equation 1 is referred to as the pressure head at locations 1 and 2, respectively. Loss refers to losses occurring in fluid flow, such as by friction. Multiplying each side of equation 1 by (ρg) yields equation 2.
P1+(ρ/2)(U1)2=P2+(ρ/2)(U2)2+Loss [2]
The (ρ/2)(U)2 term in equation 2 is used by those skilled in the art to evaluate the performance of a diffuser as shown in equation 3
CP=ΔP/(ρ/2)(U0)2 [3]
where CP is a pressure recovery coefficient, ΔP is the absolute pressure recovery or the static pressure difference between the pressure at the inlet of the diffuser and the pressure at the outlet of the diffuser, and the remaining (ρ/2)(U0)2 term is the total velocity head at the outlet of the compressor. The pressure recovery coefficient is a parameter frequently used to measure the operating performance of the diffuser. The pressure recovery coefficient is a measure of the amount of the total available velocity head at the inlet of the diffuser that is converted into static pressure.
Stated simply, the condenser inlet diffuser of the present invention not only changes the direction of flow of refrigerant vapor from a substantially vertical direction to a substantially horizontal direction with minimal flow losses, but additionally converts a portion of the kinetic energy component to a pressure head or static pressure component as shown in equation 1. That is, the condenser inlet diffuser reduces the velocity of the incoming refrigerant vapor as the refrigerant vapor flows through the inlet diffuser toward the condenser tubes while simultaneously increasing the level of static pressure. By increasing the level of static pressure, the condenser can operate at an elevated saturation temperature, thereby requiring less heat transfer surface to exchange the same amount of heat, due to higher temperature difference between the refrigerant vapor entering the condenser shell and the fluid flowing through tubes inside the condenser shell. Additionally, by reducing the velocity of the refrigerant vapor, the difference in levels of collected liquid refrigerant along a lower portion of the condenser is substantially equalized, i.e., liquid hump is minimized. Further, direct impingement of tubes of the condenser due to the flow of the refrigerant vapor is minimized.
Referring to
In addition to permitting more efficient operation of an existing refrigeration system by adding the diffuser of the present invention, while otherwise leaving the remaining system components unchanged, alternate constructions are also possible. That is, a condenser having fewer tubes than the originally installed condenser can be used, at a significant cost savings, while providing comparable operating efficiencies. The reason fewer tubes can be used is because the saturated condensing refrigerant temperature leaving the diffuser of the present invention is increased, due to the pressure recovery provided by the diffuser of the present invention, thereby providing a greater temperature gradient between the tubes and the refrigerant. For example, in one chiller configuration tested, the amount of condenser heat transfer, or tube, surface area was reduced by more than 17%, while operating more efficiently than the baseline configuration. However, it is to be understood that factors such as the size and number of the condenser tubes, types of refrigerant and secondary fluid, alternate inlet diffuser profiles, compressor discharge pipe diameter and operating loads can affect chiller system performance values.
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.
