Heat Recovery In Industrial Refrigeration

Transcription

1 The following article was published in ASHRAE Journal, August Copyright 2007 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. It is presented for educational purposes only. This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE. The search for recoverable heat from industrial refrigeration systems often begins in the engine room. Heat Recovery In Industrial Refrigeration By Douglas T. Reindl, Ph.D., P.E., Member ASHRAE; and Todd B. Jekel, Ph.D., P.E., Member ASHRAE During the past year, concepts of sustainability have received a great deal of emphasis within ASHRAE. This article explores one aspect of sustainability in the context of industrial ammonia refrigeration systems. In this article, we explore approaches of increasing the use of primary energy consumed during the course of refrigeration system operation. Specifi cally, approaches of gathering and using heat normally discarded from a refrigeration system are discussed and evaluated. Many industrial refrigeration system end-users are increasingly focusing on identifying operating principles and technologies that can effectively improve the efficiency and cost-effectiveness of their utility infrastructure. Those end-users who proactively seek opportunities to improve the energy efficiency of their operations will benefit by improved competitiveness today, but they will also be uniquely positioned to compete in a future with energy prices that are certain to escalate. Because refrigeration is often the single largest energy consumer in food production and storage facilities, it is a natural target for efficiency improvements. With the cost of energy continuing to rise, now is the time to revisit opportunities and evaluate approaches that can better use primary energy to About the Authors Douglas T. Reindl, Ph.D., P.E., is a professor and director, and Todd B. Jekel, Ph.D., P.E., is assistant director at the University of Wisconsin-Madison s Industrial Refrigeration Consortium in Madison, Wis. 22 ASHRAE Journal ashrae.org August 2007

2 achieve a facilities operational objectives. Here we explore the potential for recovering thermal energy (heat) from industrial refrigeration systems. Using recovered heat offers the potential for reducing both the direct primary energy consumption associated with refrigeration system operation, as well the consumption of primary energy used to meet heating demands directly. Potential applications for using waste heat from an industrial refrigeration system include underfloor heating (e.g., as required for freezers), cleanup water pre-heating, domestic water heating, boiler makeup water heating, and space heating (for both temperature and humidity control). As compelling as heat recovery sounds, it is possible to improperly apply and/or control heat recovery equipment; thereby, increasing the overall energy consumption of a refrigeration system relative to a base operation without heat recovery. Let s begin by first understanding the fundamentals this opportunity presents and then we will provide information to help you determine whether or not heat recovery has application in a given plant or facility. Heat Recovery Overview For the purposes of this article, we define heat recovery as gathering and using thermal energy that normally would be rejected from the system to the ambient environment. At a fundamental level, refrigeration systems absorb heat from spaces or products where it is unwanted and reject that heat at a higher energy state to the outdoor ambient environment. Heat recovery strategies attempt to find a home for a portion of the heat that is normally rejected from the system. Some key issues need to be considered in evaluating the potential for heat recovery in a given application including: Quantity of heat required, Btu (kj); Rate of heat required, Btu/h (kwt); Quality of heat required, F ( C); and Time of day the heat is required (coincidence of heat availability and demand). Theoretically, any quantity of heat that can be recovered from an industrial refrigeration system offers the potential for system efficiency improvement. Practically, the quantity of recoverable heat must be large enough to warrant the investment in capital equipment, as well as any additional operational overhead (e.g., maintenance) or increased complexity required to use the waste heat. The quality of heat is another factor that may limit the extent to which thermal energy is recoverable from an industrial refrigeration system, which is a detail that will be discussed shortly. Finally, it is desirable to have the availability of the waste heat stream be coincident with the demand. In cases where there is a time mismatch between the supply of recovered heat and demand, one or more storage tanks will be needed to accumulate the recovery-heated fluid (e.g., hot water storage). Locating Recoverable Heat For industrial refrigeration systems, the number of locations within the system suitable for recovering heat is limited. It is also important to recognize, evaluate, and reduce to the extent possible, any penalty or operational risk associated with the presence of heat recovery equipment (e.g., refrigerant-side pressure drop). The following are potential locations within an industrial refrigeration system that could supply thermal energy suitable for recovery: Oil cooling/heat rejection (fluid-cooled screw compressors) Head cooling (reciprocating compressors) High-stage discharge gas stream Reciprocating compressors Screw compressors Booster discharge gas stream (minimal opportunity) Today s screw compressors use oil for lubrication, control actuation, and rotor sealing to maintain the machine s volumetric efficiency during the compression process. A consequence of using oil in these machines is that the oil absorbs a portion of the heat of compression from the refrigerant as it moves through the compressor. As such, the oil leaving a screw compressor will be at or near the machine s refrigerant discharge temperature. Apart from the variants of liquid injection-type oil cooling strategies, maintaining proper compressor operation requires post-compression oil cooling to limit the oil s temperature rise. Thermosiphon and water-cooled (or glycol-cooled) oil cooling heat exchangers are used on screw compressor packages to lower the oil from discharge temperature to an acceptable supply temperature (e.g., 130 F [54 C]). Thus, oil cooling heat rejection is one potential source for heat recovery. Reciprocating compressors that operate at moderate to high compression ratios require a supply of water, glycol, or refrigerant to the machine s heads for cooling. This head cooling fluid (water or glycol) is a second potential source of recoverable heat recovery. Recovering heat from refrigerant gas leaving the discharge of a compressor is another readily identifiable opportunity for industrial refrigeration systems. The quality of heat available for recovery will depend on the compressor technology (reciprocating vs. screw), the compressor s operating suction pressure, discharge pressure, and load. For screw compressors, the quality of heat will also vary depending on whether the machine uses internal (i.e., liquid injection) or external oil cooling. By their very nature, compressors equipped with liquid injection oil cooling have lower discharge temperatures; therefore, a lower quality or grade of thermal energy available for recovery. Liquid-injected machines have a higher discharge mass flow rate of vapor so the quantity of heat recoverable is often greater than thermosiphon or external oil cooled packages. Let s look at how much heat we can recover and the quality of heat we have available from these sources. August 2007 ASHRAE Journal 23

3 Pressure [psia] Ammonia Condensing (~90%) 11.4 F 95 F 95 F 65 F 220 F 185 F Desuperheating (~10%) Btu/lbm-R Usable Heat (Btu/h per ton) 1,600 1,400 1,200 1, Assumes: 11.4 F ( 11.4 C) Saturated Suction Temperature 95 F (35 C) Saturated Condensing Temperature 185 F (85 C) Discharge Temperature Heat recovery rate is on a per ton of refrigeration basis Minimum Usable Temperature ( F) 190 Figure 1: Desuperheating and condensing energy split for a nonliquid-injected, high-stage twin screw compressor operating at 25 psig suction (11.4 F [ 11.4 C] saturated) and 181 psig discharge (95 F [35 C] saturated). Heat Recovery Potential Since most industrial refrigeration systems today use screw compressors, we will focus our attention on heat recovery potential from this compression technology. We will explore opportunities to recover heat from both screw compressor discharge gas and oil cooling heat exchangers. We will not consider heat recovery from reciprocating compressors (head cooling or discharge gas) here due to the limited use of this compression technology. Screw Compressor Discharge Gas In exploring the potential for heat recovery from the discharge gas stream from a screw compressor, there are four permutations of the basic options that could be considered: High-stage compressor Liquid injection oil cooled External oil cooled Booster compressor Liquid injection oil cooled External oil cooled For compressors that use liquid injection (and its variants) ,000 Enthalpy [Btu/lb m ] for oil cooling, the discharge gas temperature is suppressed to cool the oil leaving the compressor. Normally, these machines have discharge gas temperatures in the range of 130 F (54 C) whereas external oil cooled machines operate with discharge temperatures as high as 185 F (85 C). Because the discharge gas temperature from liquid-injected screw compressors is lower, this technology is less attractive for heat recovery meaning that the economic criteria for the majority of most end-users will not be met. Let s consider the more optimistic case of external oil cooled screw compressors. The discharge refrigerant gas from a compressor has ther- Figure 2: Recoverable sensible thermal energy based on minimum usable temperature. mal energy in the form of both sensible (superheat) and latent (condensing) heat. Figure 1 illustrates the compression process from suction (1) to discharge (2) for a typical fluid-cooled screw compressor operating with ammonia at 25 psig suction (11.4 F [ 11.4 C] saturated) and 181 psig condensing (95 F [35 C] saturated). In this best case scenario, only 11.5% of the total system heat rejection is available in the superheat range (higher quality) while the remaining 88.5% of the heat is available in the phase change from vapor to liquid (lower quality). Unfortunately, the highest quality of heat lies in the superheat region where the lowest quantity of heat is available. For the given operating conditions, the total heat rejection requirement to condense the discharge gas to a saturated liquid is 13.9 mbtu/h per ton of useful refrigeration (1.2 kw THERMAL / kw REFRIGERATION ). The natural question then becomes: How much thermal energy can be recovered from the compressor s discharge gas stream? The answer depends on the minimum usable temperature by the process receiving the heat. For the purposes of exploring the heat recovery potential, let s assume an ideal heat exchanger (zero approach temperature). If heat at temperatures as low as 95 F (35 C) can be utilized, the entire heat of rejection (sensible and latent) can be recovered (i.e mbtu/h ton). If higher temperatures are required, the recoverable heat will diminish as the minimum usable temperature rises as shown in Figure 2. For example, the recoverable quantity of heat at a minimum usable temperature of 110 F (43 C) is a paltry 1.3 mbtu/h ton (0.11 kw THERMAL /kw REFRIGERATION ) a portion of the sensible fraction of rejected heat from the system. An issue that arises in considering heat recovery from the discharge of external oil cooled screw compressors is the reduction in discharge temperature as the system head pressure decreases. Expect the discharge gas temperature for this type of screw compressor to decrease with lower head pressure at the rate of approximately 0.6 F/psig (4.8 C/bar). In other words, a machine that has a discharge temperature of 185 F (85 C) at discharge pressure of 180 psig (12.4 bar) can be expected 24 ASHRAE Journal ashrae.org August 2007

4 to drop to approximately 167 F (75 C) when the system head pressure drops to 150 psig (10.3 bar). The advantage of lowering system head pressure is improved efficiency of compression. This advantage significantly outweighs the decrease in recoverable heat from compressor discharge gas. A potential source for heat recovery is the discharge gas stream from booster compressors. Although recovering heat from booster discharge gas has the desirable effect of reducing high-stage compressor load, the usable heat is in a very narrow range within the superheat region. Figure 3 shows that the proportion of the total discharge superheat is higher but the fraction that is useful for heat recovery is less than a high-stage compressor. In the case being considered, discharge vapor leaves the booster at approximately 160 F (71 C) and is desuperheated to 95 F (35 C). Although there is still energy available in the superheat from 95 F (35 C) to the +10 F ( 12 C) intermediate temperature, the thermal energy is of low quality; therefore, it is usually not considered recoverable. As a fraction of the total heat rejected from the booster, only 5% is considered useful for heat recovery. While the percentage does not sound impressive, the impact of reduction of high-stage load can make this attractive if the pressure drop on the vapor side of the heat exchanger is kept below 1 psi (0.069 bar). Beyond the thermal considerations, recovering heat from the discharge gas stream of any compressor brings additional concerns including: Parasitic effects of refrigerant-side pressure drop due to presence of heat transfer equipment, which translates to increased booster horsepower; and Potential risks to refrigeration system in the event of a heat exchanger failure (e.g., water incursion directly into the refrigeration system). Recognizing that compressor discharge gas heat recovery offers a relatively modest opportunity to cost-effectively use this relatively low grade of energy, we now consider another alternative. Screw Compressor Oil Coolers Most practitioners consider thermosiphon as the state-of-theart in oil cooling for screw compressor packages because of its inherent efficiency. Thermosiphon oil cooling utilizes saturated high-pressure liquid as the coolant in a refrigerant-to-oil heat exchanger integrated into the compressor package. The lower temperature refrigerant liquid within the heat exchanger absorbs heat from the hot oil and evaporates. The high-pressure refrigerant vapor is then vented to the condenser(s) where it is recondensed and made available to continue cooling oil or meeting system loads. Thermosiphon oil cooling heat exchangers are a type of gravity flooded evaporator. Because of its principle of operation, thermosiphon oil cooling requires minimal energy for heat rejection only incremental condenser fan and pump energy. Thermosiphon oil cooling is particularly effective when properly engineered; however, it is also a technology that has Pressure [psia] 10 4 Ammonia F 95 F Figure 3: Recoverable heat in desuperheating to 95 F (35 C) with remainder as heat rejected to high stage for a booster twin screw compressor operating at 40 F ( 40 C) saturated suction and 10 F ( 12 C) saturated discharge temperature. Oil Cooler Photograph 1: Screw compressor package equipped with a glycolcooled oil cooler. dissatisfied many plant personnel regarding operation. Improperly sized thermosiphon pilot receivers, vent lines, return lines or supply lines have individually or collectively contributed to a number of problem installations. In addition, thermosiphon oil cooling systems are not particularly conducive to system expansions. When problems arise in thermosiphon systems (which happens quite frequently), those problems are difficult and frustrating to troubleshoot and solve. A less frequently applied oil cooling heat exchange technology alternative is the water-cooled (or glycol-cooled) oil cooler. In this approach, water or glycol is used as a secondary fluid to absorb heat from the oil. Photograph 1 shows a screw compressor package (booster) equipped with a glycol-cooled 160 F Non-Recoverable (~95%) 10 F Enthalpy [Btu/lb m ] Recoverable (~5%) Btu/lbm-R August 2007 ASHRAE Journal 25

5 Oil Cooling Heat Rejection (mbh) Fluid Cooled Oil Cooler Twin Screw Compressor 40 F ( 40 C) 0 F ( 18 C) 10 F ( 12 C) 30 F ( 1 C) Saturated Suction Temperature Oil Cooling Heat Rejection (mbh/ton) Fluid Cooled Oil Cooler Twin Screw Compressor Saturated Suction Temperature 40 F ( 40 C) 0 F ( 18 C) 10 F ( 12 C) 30 F ( 1 C) Saturated Condensing Temperature ( F) Saturated Condensing Temperature ( F) Figure 4: Oil cooling heat available over a range of suction and discharge conditions. oil cooler. In this case, the glycol is pumped through the oil cooler absorbing heat from the oil and then rejecting that heat to a closed-circuit fluid cooler located outdoors. It is this heat that could be recovered and used for meeting relative modest heating demands in a plant. Typically, an oil cooling heat exchanger receives hot oil from the screw compressor s oil separator at a temperature near the compressor s discharge gas temperature, which ranges between 160 F to 185 F (71 C to 85 C), and cool the oil to a supply temperature of 130 F (54 C). With oil in this operating temperature range we clearly have a heat source with reasonable quality Figure 5: Normalized oil cooling heat available over a range of suction and discharge conditions. (temperature). The question now becomes: What quantity of heat is available from an oil cooling heat exchanger? The quantity of heat available from an oil cooler will depend on a number of factors including: Size of the compressor (capacity); Operating suction pressure; Operating discharge pressure; and Part-load ratio. Figure 4 illustrates the trends in heat rejected through oil cooling heat exchangers over a range of suction and discharge/condensing conditions for a typical twin screw compressor package. SST F ( C) 10 ( 23) 0 ( 18) 10 ( 12) 20 ( 7) 30 ( 1) SDT F ( C) Discharge Temp. F ( C) Compressor Capacity Tons (kw) OCHR mbtu/h (kw) Recovered Energy Flow (gpm [Lpm]) Desuperheater Heat Exchanger Oil Cooling Heat Exchanger 95 (35) 181 (83) 419 (1474) 1102 (323) 13.5 (51) 40.1 (152) 85 (29) 176 (80) 432 (1519) 881 (258) 11.8 (33) 32.0 (121) 75 (24) 171 (77) 446 (1569) 677 (198) 10.3 (39) 24.6 (93) 95 (35) 181 (83) 538 (1892) 1056 (310) 17.2 (65) 38.4 (145) 85 (29) 175 (79) 555 (1952) 815 (239) 14.7 (56) 29.6 (112) 75 (24) 168 (76) 571 (2008) 593 (174) 12.1 (46) 21.6 (82) 95 (35) 179 (82) 682 (2399) 976 (286) 20.8 (79) 35.5 (134) 85 (29) 172 (78) 702 (2468) 716 (210) 17.3 (66) 26.0 (99) 75 (24) 163 (73) 722 (2539) 479 (140) 13.2 (50) 17.4 (66) 95 (35) 176 (80) 854 (3003) 861 (252) 24.4 (92) 31.3 (118) 85 (29) 167 (75) 878 (3088) 584 (171) 19.0 (72) 21.2 (80) 75 (24) 156 (69) 902 (3172) 340 (100) 13.0 (49) 12.4 (47) 95 (35) 171 (77) 1,058 (3721) 709 (208) 26.9 (102) 25.8 (98) 85 (29) 161 (72) 1,087 (3823) 427 (125) 19.7 (74) 15.5 (59) 75 (24) 149 (65) 1,117 (3928) 198 (58) 11.8 (44) 7.2 (27) Table 1: Comparative heat recovery potential for compressor discharge vapor versus oil cooling heat exchanger for a mid-sized high-stage screw compressor. 26 ASHRAE Journal ashrae.org August 2007

6 The availability of heat from oil cooling increases with increasing condensing temperature (pressure) and/or decreasing suction temperature (pressure). Figure 5 shows the normalized oil cooling load (oil cooling heat rejected per ton of refrigeration capacity) for the same range of discharge/condensing conditions with the same typical twin screw compressor package. For the same high stage twin screw compressor, Figure 6 shows the oil cooling load expressed as a fraction of the total heat of rejection as a function of the saturated suction temperature and saturated discharge (condensing) temperature. With fluid-cooled oil cooling heat exchangers, it is feasible to achieve recovered fluid temperatures from the oil coolers in the range of 110 F (43 C) based on a 55 F (13 C) supply. For such temperatures, the recovery fluid flow rate will depend on the fluid type (water, ethylene glycol, propylene glycol) and the oil cooling heat rejection. The recovered heat flow ranges from 3.6 gpm per 100 mbtu/h (0.465 L/min per kw) of oil cooling heat rejection (water) to 3.8 gpm per 100 mbtu/h of oil cooling heat rejection (25%wt solution of ethylene glycol in water). To place the above two heat recovery opportunities into perspective, let s look at the scale of recoverable heat using a midrange sized high-stage screw compressor with external oil cooling. Performance data for this machine over a range of saturated suction temperatures (SST) and saturated discharge temperatures (SDT) is provided in Table 1. The compressor performance data consisting of capacity, oil cooling heat rejection (OCHR) and discharge gas temperature is used to predict recovery hot water flow rates assuming an entering water temperature of 55 F (13 C) and a leaving water temperature of 110 F (43 C). The approach on each heat recovery heat exchanger is assumed to be 20 F (11 C). The estimated hot water heat recovered from both a desuperheater (located in the compressor discharge gas stream) and an oil cooling heat exchanger are provided. At low suction temperatures, the heat available for recovery from the oil cooler yields two to three times that available in the discharge gas stream. As the head pressure decreases, the heat available for recovery in both the discharge gas stream and oil cooling heat exchanger decreases, as expected, with the latter generally decreasing faster. For higher suction temperature operation, the recoverable heat from the discharge gas stream begins to approach that of the oil cooling heat exchanger during higher head pressure operation. At high suction temperatures and low condensing pressures, oil cooling loads decrease at a faster rate, which diminishes the heat available for recovery while the discharge gas stream heat recovery remains relatively flat. It is also worthwhile to note that the recoverable heat as a fraction of the compressor s refrigeration capacity decreases as the suction pressure rises. For example, the heat recovery fluid flow rate from oil cooling heat exchangers at the 10 F/95 F ( 23 C/35 C) condition is gpm per ton (0.103 L/min per kwt). When the suction temperature rises to +20 F ( 7 C) with a saturated condensing temperature of 95 F (35 C), the fluid flow rate to recover heat from the oil cooling heat exchanger drops threefold to gpm per ton ( L/min per kwt). Finally, the results shown here for heat recovery from the discharge gas stream do not include any effects of refrigerant-side pressure drop on the performance of compressors operating. The majority of the operating cost savings with the above heat recovery options is attributable to a reduction in primary energy (natural gas or propane) that would normally be used for heating. Figure 7 shows the displaced heating energy cost (savings) due to heat recovery for a range of primary fuel costs and flow rates. The recovered heating energy is assumed to offset the operation of an 80% efficient boiler that would normally be used to provide the required heat. Recovering heat from an industrial refrigeration system has the added benefit Advertisement formerly in this space. August 2007 ASHRAE Journal 27

7 Fraction of Total Heat Rejected [ ] 0.35 Saturated Suction Temperature 0.30 Fluid Cooled Oil Cooler Twin Screw Compressor F ( 40 C) 0 F ( 18 C) 10 F ( 12 C) 30 F ( 1 C) Saturated Condensing Temperature ( F) 28 F ( 33 C) Heat Cost Savings ($/yr) 160, , , ,000 80,000 60,000 40,000 20, $1.00/therm $0.75/therm $0.50/therm $0.25/therm Average Water Flow (gpm) 50 Figure 6: Fraction of total heat of rejection that appears as an oil cooling load over a range of suction and discharge conditions. Figure 7: Annual heating cost savings for a 55 F (31 C) range assuming the offset operation of an 80% efficient boiler. of reducing the heat rejection load on the evaporative condensers, which leverages the primary heating energy savings. It is reasonable to assume that a plant will realize some incremental savings in evaporative condenser fan energy with the load reduction on the condensers. Conservatively, the added pump energy for the heat recovery system offsets any reduction in evaporative condenser water pump energy. Other Considerations In evaluating the potential for heat recovery in specific plant applications, a number of other factors warrant careful consideration. As mentioned previously, the demand for recovered heat needs to be compared with the available supply. In situations where there is a mismatch, some form of heat storage will be needed. Related to this issue is providing redundancy in heat rejection. If you are considering recovering heat from oil cooling heat exchangers, it is important to provide some form of heat rejection redundancy such as the installation of one or more closed-circuit fluid coolers that can be used during periods when there is no plant demand for recovered heat. With a secondary fluid such as glycol in the heat recovery loop, another heat exchanger will be required if the application calls for heating potable water. Be sure to check the mechanical code in the application s jurisdiction as the heat exchanger likely will need to be a double-walled design. Finally, for retrofits, be sure to work with your compressor manufacturer to evaluate any special requirements for oil circulation, controls, interlocks, etc., on the package. For desuperheaters, it is important to evaluate the operational risks associated with heat exchanger failure. Refrigerant-side pressure drop needs to be minimized to make this option viable. Finally, implementation of a refrigerant vapor bypass should be considered for reliability. Conclusions As the cost of fuels used for heating increases, the opportunities to economically recover heat from a refrigeration system grow. Traditional heat recovery approaches focus on recovering heat from the high stage compressor discharge gas stream (desuperheater). An alternative that should not be overlooked is recovering heat from oil cooling heat exchangers on screw compressor packages. Advantages of this option include: Ability to recover heat from both high-stage and booster compressor oil coolers; Allows recovery of relatively high-quality waste heat; Results in an oil cooling approach that is less prone to problems compared to thermosiphon oil cooling; Use of sensible heat transfer fluid and pump provides more control of the oil cooling (taking Mother Nature out of the traditional thermosiphon loop); Accommodates addition of compressors during expansions more easily than a thermosiphon oil cooling system; Great retrofit opportunity for liquid injection oil cooled compressors; Makes compressor oil cooling independent of the system pressures; therefore, facilitates easier startup; and Allows the use of plate-frame heat exchangers without worrying about refrigerant-side pressure drop. Although the advantages are compelling, other factors need to be considered in implementing this heat recovery approach. First, each compressor package needs to be fitted (or retrofitted) with an appropriate oil cooling heat exchanger. Second, the demand for heat needs to be matched with the production of heat by the oil coolers. If there is a mismatch between demand and supply, extra infrastructure is needed to accommodate the mismatch. Finally, the economics of recovering heat from oil cooling heat exchangers need to be evaluated on a case-by-case basis. 28 ASHRAE Journal ashrae.org August 2007

8 Advertisement formerly in this space.